Vapor Compression Evaporative Air Conditioning systems and Components

ABSTRACT

Novel vapor compression evaporative cooling systems which use water as a refrigerant are provided, as are methods for using same. Also provided are novel compressors, compressor components, and means for removing noncondensibles useful in such cooling systems.

This application is a divisional application of U.S. patent applicationSer. No. 10/768,908, filed Feb. 2, 2004; which is a divisionalapplication of U.S. patent application Ser. No. 09/964,401, filed Sep.28, 2001, now issued as Patent No. U.S. Pat. No. 6,684,658 B2; which isa divisional application of U.S. patent application Ser. No. 09/126,325,filed Jul. 31, 1998, now issued as Patent No. U.S. Pat. No. 6,427,453B1; the disclosures of which are incorporated herein by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates to vapor-compression evaporative cooling systemsthat use water as a refrigerant in an open system, and in particular, tovapor-compression evaporative cooling systems capable of processinglarge volumetric flow rates of water vapor and removing noncondensiblesfrom the system and to methods using such systems. This invention alsorelates to low-friction, positive-displacement compressors useful insuch cooling systems and to means for removing noncondensibles from suchcooling systems.

2. Description of the Background

Conventional vapor-compression air conditioning systems employ a workingfluid such as chlorofluorocarbons (CFCs). Liquid CFC is introduced intoa low-pressure heat exchanger where it absorbs heat at a low temperatureand vaporizes. A compressor repressurizes the vapors that are introducedto a high-pressure heat exchanger where heat is rejected to theenvironment and the vapors condense. The condensate is reintroduced intothe low-pressure heat exchanger, thus completing the cycle.

The use of CFCs raises two important environmental concerns. First, CFCsare stable enough to enter the stratosphere where they decompose tochlorine free radicals that catalyze the destruction of ozone. This isunfortunate because ozone absorbs ultraviolet radiation which damagesDNA in plants and animals. Second, CFCs absorb infrared radiation whichcontributes to global warming.

Because CFCs cannot be released into the environment, they must becontained within the air conditioning system. The evaporator andcondenser heat exchangers have a sizable temperature difference betweenthe ambient environment and the working fluid (about 10 to 15° C.) whichgreatly reduces the Carnot efficiency. Further limiting the efficiencyis the fact that the condenser rejects heat at the dry-bulb temperature.The wet-bulb temperature is generally about 5-30° C. less than thedry-bulb temperature. Thus, if heat were rejected at the wet-bulbtemperature, the Carnot efficiency could be improved even more.

In addition, compressors used in conventional systems typically havecompressing components that are in direct contact with each other. Theclose fit between components has heretofore been necessary to preventblow-by of high-pressure compressed vapors. However, the frictionresulting from the close contact between components reduces efficiency,creates heat and causes wear on the components.

Although the use of water in place of CFCs as the air-conditioningworking fluid has been considered, proposed systems have been generallyunworkable because the vapor density is very low requiring large volumesof water vapor to be compressed.

One study by the Thermal Storage Applications Research Center of theUniversity of Wisconsin, The Use of Water as a Refrigerant, Report No.TSARC 92-1, March 1992, studied the use of water as a refrigerant. Thisstudy concluded that for water-based air conditioning, positivedisplacement compressors are not suitable for use in such systems.Rather, only dynamic compressors are suitable.

Although “swamp cooler” air conditioners are employed in arid regions ofthe United States that have low wet-bulb temperatures, they have limitedusefulness. In swamp coolers, ambient air is contacted with water whichevaporates and cools the air. No external power is required other thanfor air-handling blowers. Unfortunately, these simple devices arerestricted to regions of low humidity (e.g., Arizona, New Mexico) andare not suitable for many regions of the world. Further, although theair is cooler, it has increased humidity which can make the air feel“clammy.”

SUMMARY OF THE INVENTION

There is therefore a need for an environmentally friendly, efficient andeconomical means for air conditioning in all types of climates. Thepresent invention overcomes the above noted deficiencies in the art byproviding air conditioning systems that use water as the working fluidrather than CFCs, thus eliminating potential CFC emissions. Thesesystems are not limited to regions of low humidity. The presentinvention is directed to cooling systems that are 1.7 to 3.9 times moreefficient than conventional air conditioning systems and that havemanufacturing costs less than, or competitive with, conventional airconditioning systems.

In addition, unlike the teachings of the literature, it has beendiscovered that high-volume, low-pressure positive displacementcompressors can be utilized in cooling systems that use water as theworking fluid. It has further been discovered that because of therelatively low pressures (i.e., 0.2-0.7 psia) in the compressors of thecooling systems of the present invention, the gaps between thecompressing components can be comparatively large, and that such largegaps are not only acceptable, but actually can be beneficial from bothan efficiency and wear standpoint. Because of the low friction, thenovel compressors can be scaled up to the necessary size. For example,such a gap-containing, positive displacement compressor can process the1400 ft³/min of low-pressure water vapor needed to produce 3 tons ofcooling.

In addition, it has been discovered that water, with or without suitablewicking material, can be used to fill the gaps between the components,and thereby create an effective, but low-friction seal between thecompressing components. Thus, the present invention is also directed tonovel positive displacement compressors which are useful in airconditioning systems using water as the working fluid. These compressorsinclude novel compressors which are useful in the disclosed systems aswell as in other applications. The present invention is also directed tonovel pumps useful for removing noncondensibles from the disclosedcooling systems as well as in other applications. Finally, the presentinvention is directed to novel seals and mounting apparatus useful inthe disclosed compressors.

In accordance with one embodiment of the present invention, avapor-compression evaporative air conditioning system is provided thatcomprises: an evaporator; a room air contactor for directly exchangingheat between room air and a quantity of water from the evaporator; meansfor compressing a volume of water vapor, thereby creating a vacuum onthe water in the evaporator, the means for compressing comprising apositive displacement compressor, the compressor comprising an inlet andan outlet, wherein low-pressure water vapors from the evaporator enterthe inlet and compressed water vapors exit the outlet; a condenser forreceiving the compressed water vapors; means for reducing a watercontent of the vapors exiting the condenser; means for removingnoncondensibles from the condenser; and an ambient air contactor fordirectly exchanging heat between the ambient air and water from thecondenser. The positive displacement compressor is preferably alow-friction compressor comprising at least two compressing components,which do not substantially contact one another. The advantages of thissystem include that it is an efficient low-friction system capable offunctioning in humid environments.

The compressing components may comprise: an inner gerotor, an outergerotor and a housing; an orbiting (or mobile) scroll, a stationary (orfixed) scroll and a housing; a housing and a piston; a housing, a rotor,and a flap; an inner drum, an outer drum and a swinging vane; or ahousing, a rotor and a sliding vane. In a preferred embodiment there isa gap between at least two of the compressing components. Water or waterand a wick may be used as a sealant in the gap.

In one embodiment of this system, the means for compressing water vaporcomprises a gerotor compressor comprising an inner gerotor and an outergerotor, the inner gerotor disposed within the outer gerotor, eachgerotor comprising a plurality of teeth. The inner gerotor has one lesstooth than the outer gerotor, thereby creating a void volume between theinner gerotor and the outer gerotor. An inlet port and a discharge portcommunicate with the void volume. The discharge port may have a variableport mechanism that changes the position of a leading edge of thedischarge port. This variable port mechanism may be positioned usingelectrically actuated means controlled by a thermocouple signal.

The variable port mechanism may comprise an electrically controlledservo motor, the motor rotating a threaded rod, a bellows, and anon-rotating nut coupled to the bellows, the rod axially positioning thenon-rotating nut. Alternatively, the variable port mechanism maycomprise a plurality of plates disposed adjacent to the discharge portand means for sequentially moving the plates to vary the leading edge ofthe discharge port. The variable port mechanism may be positioned usinga bellows, actuated by a bulb containing a liquid, wherein the liquid inthe bulb has a vapor pressure proportional to the condenser temperaturewhich acts on the bellows.

The gerotor compressor may further comprise an electric motor fordriving the gerotor compressor, a first pump for pumping cooled waterfrom the evaporator to a packing in the room air contactor, a filterdisposed between the room air contactor and the evaporator, whereinwater from the room air contactor flows through the filter to theevaporator, a second pump for pumping water from the condenser to apacking in the ambient air contactor, and a fan for driving ambient aircountercurrently against the packing.

Because of the low friction between the compressing components of thecompressors of the present invention, the compressors of the presentinvention use novel actuation means to actuate the gerotors.

For instance, one embodiment uses a low-friction gerotor compressor inwhich a first drive shaft drives the outer gerotor, and the actuationmeans comprises an internal gearbox containing a plurality of spurgears, the plurality being an odd number. One of the spur gears iscoupled to the first drive shaft and another of the spur gears iscoupled to a second drive shaft, the second drive shaft being offsetfrom the first drive shaft, thereby suspending the gearbox between thefirst drive shaft and the second drive shaft. The first drive shaft iscoupled to the outer gerotor through a plate that comprises a pluralityof prongs in contact with a plurality of holes in the outer gerotor. Thesecond drive shaft is coupled to the inner gerotor.

In another embodiment, a different novel actuated gerotor compressor isused. In this compressor, a first shaft drives the outer gerotor and theactuation means comprises a spur gear set comprised of a large gearcoupled to the outer gerotor, the large gear containing a plurality ofteeth on an inside diameter, and a small gear coupled to the innergerotor, the small gear containing a plurality of teeth on an outsidediameter, the large gear meshing with the small gear, and furthercomprised of a second shaft about which the inner gerotor spins, whereinthe second shaft contains a crook establishing an offset between thefirst shaft and the second shaft. Preferably, for cooling andlubrication purposes, the gears are immersed in liquid water. A gear setmay be attached to a bottom portion of the inner gerotor allowing forpower take off.

In an alternative embodiment using still another novel actuated gerotorcompressor, the actuation means may comprise a plurality of rollersattached to the inner gerotor, wherein the rollers extend beyond aplurality of walls of the inner gerotor and are in contact with theouter gerotor, and wherein the outer gerotor drives the inner gerotorthrough the rollers. In this embodiment, the inner gerotor may bemounted on a rotating shaft and the rotating shaft extends outside ofthe compressor housing.

In still another embodiment using a novel actuated gerotor compressor,the actuation means comprises a large gear coupled to the outer gerotor,the large gear comprising a plurality of teeth on an inside diameter, asmall gear coupled to the inner gerotor, the small gear comprising aplurality of teeth on an outside diameter, the large gear meshing withthe small gear, and a stationary central shaft, wherein the stationarycentral shaft contains two crooks that create an offset between an axisof the inner gerotor and an axis of the outer gerotor, and wherein thestationary shaft comprises a first end and a second end, the first endof the stationary shaft affixed to a first perforated housing end platethrough a pivotable mount that prevents rotation of the stationary shaftand the second end of the stationary shaft located in a rotating bearingcup coupled to the outer gerotor. In this embodiment, the gerotorcompressor may further comprise a second perforated housing plate, afirst perforated rotating plate and a second perforated rotating plate,such that both the rotating plates are connected to the outer gerotor,and a first stationary plate and a second stationary plate adjacent toboth gerotors, the first stationary plate containing an inlet port andthe second stationary plate containing a discharge port. Alternatively,the inlet and outlet port can be placed in one of the plates.Preferably, the gears are immersed in liquid water to provide coolingand lubrication.

In the novel air conditioning system disclosed herein, the system mayfurther comprise means for inhibiting microorganisms in the water in theroom air contactor, such as an ozone generator or UV radiation. Inaddition, the means for removing noncondensibles may comprise anaspirator or a vacuum pump, such as the novel pumps disclosed below.

In other embodiments of the disclosed system, the compressor means maycomprise a novel low-friction scroll compressor.

In yet another embodiment of the disclosed system, the compressor meanscomprises a novel actuated flap compressor. This compressor comprises: acompressor housing, the housing having an interior wall, an inlet, andan outlet; a rotor disposed in the housing; a flap, the flap having afirst end and a second end, the first end being coupled to the rotor andthe second end being propelled in an outward direction during rotationof the rotor; and means for preventing the second end of the flap fromtouching the interior wall of the housing.

In still another embodiment, the compressor means comprises a novelmulti-vane actuated flap compressor. This compressor preferablycomprises: an outer drum having an axis; an inner drum rotatablydisposed in the outer drum; a plurality of vanes, each vane having afirst end and a second end opposite the first end, the vanes pivotallyattached to the inner drum at the first end and having a vane tip at thesecond end, the vane tips being propelled radially outward duringrotation of the inner drum; a connecting rod coupled to each vane tip,the rods maintaining a gap between the vane tips and the outer drum; andcoupling means for causing the connecting rods to rotate about the axisof the outer drum.

Alternatively, the compressor means may be a novel low-frictionreciprocating compressor comprising: a compressor housing; anoscillating center shaft disposed partly within the housing, the shaftcomprising a top end and a bottom end, the top end comprising aprotrusion which rides in a sinusoidal groove in a rotating cam drivenby a motor; and at least one plate disposed in the housing and attachedto the shaft and oscillating therewith, the at least one plate having agroove through which water flows to make a seal between the compressorhousing and the plates. In one embodiment of the reciprocatingcompressor, the cam contains a plurality of sinusoidal grooves.

In the novel air conditioning systems disclosed herein, the componentsmay be disposed in three concentric chambers. In one such embodiment theambient air contactor is disposed in an outermost chamber of theconcentric chambers, the compressor means and the evaporator aredisposed in an innermost chamber of the concentric chambers, and thecondenser is disposed in a middle concentric chamber. In another system,comprising two concentric chambers, the ambient air contactor isdisposed in an outermost chamber of the concentric chambers, and thecompressor means, the evaporator and the condenser are disposed in aninnermost concentric chamber.

The novel systems disclosed herein may further comprise means forproviding make-up water to the evaporator and condenser, which ispreferably accomplished using one or more float valves. In addition, theroom air contactor may comprise a spray tower to place water from theevaporator in direct contact with the room air. The room air contactormay comprise a packing, such that the water from the evaporator passesover the packing, and the room air passes through the packing. Thepacking preferably comprises corrugated chlorinated polyvinyl chloride.In the disclosed embodiments, the condenser may be a spray condenser,jet condenser, or may comprise a packing.

The present invention is also directed to a novel method for cooling aircomprising the steps of: compressing a large volume of low-pressurewater vapor with a compressor, thereby creating a vacuum on a quantityof water in an evaporator and causing evaporation and the water to becooled; pumping cooled water from the evaporator and contacting thecooled water countercurrently with room air in a room air contactor,thereby cooling room air; routing water from the room air contactor tothe evaporator, causing the water to flash and cool; sending compressedwater vapors exiting the compressor to a condenser for condensation;countercurrently directly contacting the water vapors exiting thecondenser with a stream of chilled water from the evaporator to reducethe water content from air; removing noncondensibles from the condenser;routing liquid water from the condenser to an ambient air contactor,where ambient air is contacted countercurrently with liquid water pumpedfrom the condenser; providing make-up water to replace evaporated water;and draining salt water.

Preferably, the compressor is a positive displacement compressor. Morepreferably, the compressor is a low-friction positive displacementcompressor comprising at least two compressing components, in which thecompressing components do not substantially contact each other, i.e.,although some contact can occur without departing from the spirit andscope of the invention, generally there are clearance gaps, whichpreferably may be a few thousandths of an inch, between components. Themethod may further comprise the step of spraying water into thecompressor to prevent temperature increase during the compression stage.

In one embodiment of the method, water from the room air contactor mayflow countercurrently through a plurality of evaporators. Alternately,condensation may occur in multiple stages. In still another embodimentof the invention, both evaporation and condensation take place inmultiple stages. Noncondensibles may be removed by one or a plurality ofcompressors.

The present invention is also directed to novel methods of cooling airusing multistage systems. One such method comprises the steps of:compressing a large volume of low-pressure water vapor in a plurality ofcompressor stages, thereby creating a vacuum on a quantity of water in aplurality of evaporators and causing the water to be cooled; pumpingcooled water from the evaporators and contacting the cooled watercountercurrently with room air in a room air contactor, thereby coolingroom air; routing water from the room air contactor to the evaporators,causing the water to flash and cool; sending compressed water vaporsexiting the last compressor stage to a condenser for condensation;countercurrently directly contacting the water vapors exiting thecondenser with a stream of chilled water from at least one of theevaporators to reduce the water content from air; removingnoncondensibles from the condenser; routing liquid from the condenser toan ambient air contactor, wherein ambient air is contactedcountercurrently with liquid water pumped from the condenser; providingmake-up water to replace evaporated water; and draining salt water.Condensation may take place in a single stage or in multiple stages. Thecompressor stages preferably comprise one or more positive displacementcompressors or one or more dynamic compressors. However, in themultistage systems disclosed herein, the compressor stages may be eitherpositive displacement compressors, or dynamic compressors, or a mixtureof each.

The present invention is also directed to novel low-friction positivedisplacement compressors useful in the cooling systems of the presentinvention as well as in other applications. They have the advantage oflow friction and high efficiency. These compressors comprise at leasttwo compressing components, such that the compressing components do notsubstantially contact one another. The compressing components maycomprise: an inner gerotor, an outer gerotor and a housing; an orbitingscroll, a stationary scroll and a housing; a housing and a piston; ahousing, a rotor and a sliding vane; a housing, a rotor and a flap; oran inner drum, an outer drum and a swinging vane, and there is a gapbetween at least two of the compressing components. Water, or water anda wick may be used as a sealant in the gap.

One such novel compressor comprises a gerotor compressor comprising aninner gerotor and an outer gerotor, the inner gerotor disposed withinthe outer gerotor, each gerotor comprising a plurality of teeth. Theinner gerotor has one less tooth than the outer gerotor, therebycreating a void volume between the inner gerotor and the outer gerotor.In addition, there is a gap between the inner gerotor and outer gerotor.The gerotor compressor further comprises an inlet port and a dischargeport; the ports communicate with the void volume.

The discharge port may have a variable port mechanism that changes theposition of a leading edge of the discharge port. In one embodiment, thevariable port mechanism comprises an electrically controlled servomotor, the motor rotating a threaded rod, a bellows, and a non-rotatingnut coupled to the bellows, the rod axially positioning the non-rotatingnut. The variable port mechanism may be positioned using electricallyactuated means. In another embodiment, it may be positioned using abellows, the bellows being actuated by a bulb containing a liquid, theliquid in the bulb having a vapor pressure proportional to the condensertemperature which acts on the bellows. In yet another embodiment, thevariable port mechanism comprises a plurality of plates disposedadjacent to the discharge port and means for sequentially moving theplates to vary the leading edge of the discharge port.

The present invention is also directed to novel low-friction gerotorcompressors which use actuation means to actuate the gerotors allowingfor reduced friction. In one such embodiment, a first drive shaft drivesthe outer gerotor and the actuation means comprises an internal gearboxcontaining a plurality of spur gears, the plurality being an odd number,and wherein one of the spur gears is coupled to the first drive shaftand another of the spur gears is coupled to a second drive shaft, thesecond drive shaft being offset from the first drive shaft, therebysuspending the gearbox between the first drive shaft and the seconddrive shaft, and the first drive shaft is coupled to the outer gerotorthrough a plate that comprises a plurality of prongs in contact with aplurality of holes in the outer gerotor. A second drive shaft is coupledto the inner gerotor.

In another novel actuated gerotor compressor, a first drive shaft drivesthe outer gerotor and the actuation means comprises a spur gear setcomprised of a large gear coupled to the outer gerotor, the large gearcontaining a plurality of teeth on an inside diameter, and a small gearcoupled to the inner gerotor, the small gear containing a plurality ofteeth on an outside diameter. In this embodiment the large gear mesheswith the small gear, and there is a second shaft about which the innergerotor spins. This second shaft contains a crook establishing an offsetbetween the first shaft and the second shaft.

In another embodiment of a novel actuated gerotor compressor, theactuation means comprises a plurality of rollers attached to the innergerotor, wherein the rollers extend beyond a plurality of walls of theinner gerotor and are in contact with the outer gerotor, and the outergerotor drives the inner gerotor through the rollers.

In another embodiment, the inner gerotor and outer gerotor are disposedin a housing, a first drive shaft drives the outer gerotor, and theactuation means comprises a spur gear set comprised of a large gear,coupled to the outer gerotor, the large gear containing a plurality ofteeth on an inside diameter, and a small gear coupled to the innergerotor, the small gear containing a plurality of teeth on an outsidediameter. In this embodiment, the large gear meshes with the small gear,and there is a second shaft attached to the inner gerotor which spins ona bearing means, such as bearings affixed to the housing.

In still another embodiment, a first drive shaft drives the innergerotor, and the actuation means comprises a spur gear set comprised ofa large gear coupled to the outer gerotor, the large gear containing aplurality of teeth on an inside diameter, and a small gear coupled tothe inner gerotor, the small gear containing a plurality of teeth on anoutside diameter, wherein the large gear meshes with the small gear, andfurther comprises a second nonrotating shaft about which the outergerotor spins, wherein the second shaft contains a crook establishing anoffset between the first and the second shafts.

In still another embodiment, the actuation means comprises a large gearcoupled to the outer gerotor, the large gear comprising a plurality ofteeth on an inside diameter, a small gear coupled to the inner gerotor,the small gear comprising a plurality of teeth on an outside diameter,the large gear meshing with the small gear, and a stationary centralshaft, wherein the stationary central shaft contains two crooks thatcreate an offset between an axis of the inner gerotor and an axis of theouter gerotor, and wherein the stationary shaft comprises a first endand a second end, the first end of the stationary shaft affixed to afirst perforated housing end plate through a pivotable mount thatprevents rotation of the stationary shaft and the second end of thestationary shaft located in a rotating bearing cup coupled to the outergerotor. Preferably, the pivotable mount prevents the stationary centralshaft from rotating, but allows for angular and axial variation.

In this embodiment, the pivotable mount may comprise a ring, spokes anda hub, which are coupled to the shaft. The ring has a spherical outerdiameter which is disposed within an inlet of the first perforatedhousing end plate. In addition, the gerotor compressor may furthercomprise a second perforated housing plate, a first perforated rotatingplate and a second perforated rotating plate, wherein both the rotatingplates are connected to the outer gerotor, and a first stationary plateand a second stationary plate which are adjacent to the inner and outergerotors, the first stationary plate containing an inlet port and thesecond stationary plate containing a discharge port.

The present invention is also directed to novel low-friction scrollcompressors. One such compressor comprises a stationary scroll havingflutes and an orbiting scroll having flutes, the orbiting scrollorbiting around the stationary scroll. The flutes of the scrolls areseparated by a gap.

The scroll compressor of the present invention may have novel means forcreating orbiting motion. This compressor comprises a stationary scroll,an orbiting scroll, and means for causing the orbiting scroll to orbitaround the stationary scroll, the means comprising a first gear affixedto the stationary scroll, an orbiting arm affixed to the first gear, asecond intermediary gear attached to the orbiting arm, and a third gearattatched to the orbiting scroll, wherein the second intermediary geardrives the third gear.

Still other embodiments of the invention are directed to novel slidingvane compressors which comprise a rotor, a sliding vane and a housing,and means for reducing friction between the vane, the rotor and thehousing. In one such embodiment, the compressor comprises: a compressorhousing, the housing having an interior wall, an inlet, and an outlet; arotor disposed in the housing; a flap, the flap having a first end and asecond end, the first end being coupled to the rotor and the second endbeing propelled in an outward direction during rotation of the rotor;and means for preventing the second end of the flap from touching theinterior wall of the housing.

In still another embodiment, a novel multi-vane compressor comprises: anouter drum having an axis; an inner drum rotatably disposed in the outerdrum; a plurality of vanes, each vane having a first end and a secondend opposite the first end, the vanes pivotably attached to the innerdrum at the first end and having a vane tip at the second end, the vanetips being propelled radially outward during rotation of the inner drum;a connecting rod coupled to each vane tip, the rods maintaining a gapbetween the vane tips and the outer drum; and coupling means for causingthe connecting rods to rotate about the axis of the outer drum. In thisembodiment, the inner drum is preferably rotatably driven by a firstshaft, and the coupling means comprises an offset shaft to which theconnecting rod is coupled, the offset shaft being coaxial with the axisof the outer drum; and a torque coupler for transmitting rotationalforce to the offset shaft. Preferably, water is used as a sealant in thegaps.

Still another embodiment is directed to a novel low-frictionreciprocating compressor, comprising: a compressor housing; anoscillating center shaft disposed partly within the housing, the shaftcomprising a top end and a bottom end; and at least one plate disposedin the housing and attached to the shaft and oscillating therewith, theat least one plate having a groove through which water flows to make aseal between the compressor housing and the plates. In a preferredembodiment, the top end of the shaft has a protrusion which rides in asinusoidal groove in a rotating cam driven by a motor. Alternately, thecam may contain a plurality of sinusoidal grooves.

The present invention is also directed to novel pumps useful in removingnoncondensibles. Possible methods for purging noncondensiblesinclude: 1) periodically flooding the condenser with liquid water topush out accumulated noncondensibles, 2) employing an aspirator in whichthe vacuum at the throat of the venturi draws out noncondensibles, and3) employing a mechanical vacuum pump. One such embodiment comprises avacuum pump which comprises a cylinder, a piston disposed in thecylinder, an inlet valve disposed in the cylinder, a sprayer that drawswater into the cylinder, and a vent disposed in the cylinder fordischarging noncondensibles and excess water. The vacuum pump is drivenby a gear mounted on a main drive shaft, the gear connected to aplurality of reduction gears, wherein a first cam surface and a secondcam surface are mounted on one of the reduction gears, a first rollerrides on the first cam surface and a second roller rides on the secondcam surface, and the first roller drives the piston and the secondroller drives the inlet valve.

Another novel vacuum pump comprises a cylinder, piston disposed in thecylinder, a crank, a check valve disposed in the cylinder, and means forspraying water into the cylinder of the vacuum pump, wherein the pistonis driven by the crank in a first and a second direction opposite thefirst direction, the piston comprising a first end, a second end, aplurality of notches, a plurality of perforations extending from thefirst end to the second end, and a flexible flap attached to the secondend of the piston and covering one or more of the perforations, whereinthe flap opens when the piston moves in the first direction and closeswhen the piston moves in the second direction.

Still another novel vacuum pump comprises: a first column and a secondcolumn, the columns being partially filled with liquid and having avapor space; means for causing the liquid to oscillate in the columns;inlet means for allowing uncompressed gas to enter each of the columns;outlet means for discharging compressed gas from each of the columns;and means for spraying a fine mist of liquid into the vapor space of thefirst and the second columns. Preferably, the means for causingoscillation comprises a chamber connecting the first and the secondcolumns and a reciprocating piston disposed in the chamber. The outletmeans for each column preferably comprises a check valve. Thisoscillating pump has the ability to isothermally compress a mixture ofnoncondensible and condensible gases to a very high compression ratio.

Another novel vacuum pump is a gerotor vacuum pump comprising an outergerotor and a center gerotor disposed within the outer gerotor, whereinthe center gerotor is mounted on a main drive shaft and the outergerotor is positioned by a plurality of guide rollers. Alternatively,the center gerotor is mounted on a main drive shaft and the outergerotor is mounted within a single ball bearing.

The volumetric load on the aspirator or vacuum pump can be greatlyreduced by condensing most of the water and increasing the partialpressure of the noncondensibles. The present invention employs a novelmethod for removing water vapor from noncondensibles in a stream of airand water vapor comprising passing the stream through a packed columnwith chilled water flowing countercurrently. Preferably, the packedcolumn comprises structured packing (e.g., corrugated polyvinylchloride) or dumped packing (e.g., ceramic saddles).

Still another embodiment is directed to a novel pivotable mountingapparatus for mounting a stationary shaft to a housing, which preventsrotation of the shaft, but allows for angular and axial variation. Thisapparatus comprises a ring, spokes and hub, coupled to the shaft. Thering has a spherical outer diameter, which is disposed within acylindrical shaped opening in the housing.

Still another embodiment is directed to a novel low-friction rotaryshaft seal comprising: a journal for receiving a rotary shaft, thejournal configured to create a gap between the shaft and the journal,the journal further comprising a journal face; means for supplying waterto the gap; and a bellows seal, the seal resting on the journal facewhen the shaft is stationary and lifting off the face when the shaftrotates.

DESCRIPTION OF THE DRAWINGS

FIG. 1 Schematic of vapor compression evaporative cooler 100.

FIG. 2 Schematic of vapor compression evaporative cooler 101.

FIG. 3 Depiction of the coefficient of performance of cooler 101 undervarious conditions.

FIG. 4 Depiction of the coefficient of performance of R-12 vaporcompression refrigeration.

FIG. 5 Schematic cross sectional view of compressor 230 incorporatedinto cooler 200.

FIG. 6 (a-f) Schematic top views of sliding vane compressor 3300 indifferent stages of its rotational cycle.

FIG. 7 Three dimensional schematic of sliding vane compressor 3300.

FIG. 8 Exploded side view of rotor 3302 of compressor 3300.

FIG. 9 Schematic cross sectional view of groove 3342 detail of rotor ofcompressor 3300.

FIG. 10 (a) Top view of slip ring 3326 of rotor 3302 of compressor 3300;(b) top view of counterweight 3328 of rotor 3302; (c) top view of topcap 3320 of rotor 3302; and (d) bottom view of bottom cap 3330 of rotor3302.

FIG. 11 (a) Side view of sliding vane 3351 of compressor 3300; (b) crosssectional view of sliding vane 3351 taken along plane A-A of (a); and(c) cross sectional view of sliding vane 3351 taken along plane B-B of(a).

FIG. 12 Top view of compressor housing 3301 of compressor 3300.

FIG. 13 Side view of compressor housing 3301 taken along plane A-A ofFIG. 12.

FIG. 14 (a) Top view of top housing end plate 3303 of compressor 3300;and (b) side view of top housing end plate 3303 taken along plane A-A of(a).

FIG. 15 (a) Top view of bottom housing end plate 3305 of compressor3300; and (b) side view of bottom housing end plate 3305 taken alongplane B-B of (a).

FIG. 16 Schematic cross sectional view of compressor 3300 incorporatedinto cooler 3200.

FIG. 17 (a-f) Schematic top views of sliding vane compressor 4300 indifferent stages of its rotational cycle.

FIG. 18 Three dimensional schematic view of compressor 4300.

FIG. 19 Exploded side view of rotor 4302 of compressor 4300.

FIG. 20 Perspective view of sidewall 4340 and sliding vane 4308 ofcompressor 4300.

FIG. 21 Top view of compressor housing 4301 of compressor 4300.

FIG. 22 Side view of compressor housing 4301 taken along plane A-A ofFIG. 21.

FIG. 23 (a) Top view of top housing end plate 4303 of compressor 4300;and (b) side view of top housing end plate 4303 taken along plane A-A of(a).

FIG. 24 (a) Top view of bottom housing end plate 4305 of compressor4300; and (b) side view of bottom housing end plate 4305 taken alongplane B-B of (a).

FIG. 25 (a-f) Schematic top views of actuated sliding vane compressor5300 in different stages of its rotational cycle.

FIG. 26 Exploded side view of rotor 5302 of compressor 5300.

FIG. 27 (a) Top view of slip ring 5326 of rotor 5302; (b) top view oftop cap 5320 of rotor 5302; and (c) bottom view of bottom cap 5330 ofrotor 5302.

FIG. 28 Perspective view of sidewall 5340 and sliding vane 5308 ofcompressor 5300.

FIG. 29 (a) Side view of sliding vane 5308 of compressor 5300; (b) crosssectional view of sliding vane 5308 taken along plane A-A of (a); and(c) cross sectional view of sliding vane 5308 taken along plane B-B of(a).

FIG. 30 (a) Top view of top housing end plate 5303 of compressor 5300;and (b) side view of top housing end plate 5303 taken along plane A-A of(a).

FIG. 31 (a) Top view of bottom housing end plate 5305 of compressor5300; and (b) side view of bottom housing end plate 5305 taken alongplane B-B of (a).

FIG. 32 (a-f) Schematic top views of actuated flap compressor 6300 indifferent stages of its rotational cycle.

FIG. 33 Three dimensional schematic view of actuated flap compressor6300.

FIG. 34 Exploded side view of rotor 6302 of compressor 6300.

FIG. 35 (a) Top view of slip ring 6326 of rotor 6302; (b) top view oftop cap 6320 of rotor 6302; and (c) bottom view of bottom cap 6330 ofrotor 6302.

FIG. 36 Perspective view of sidewall 6340 and flap 6308 of rotor 6302 ofcompressor 6300.

FIG. 37 Side view of flap 6308.

FIG. 38 Top view of flap 6308.

FIG. 39 Top view of compressor housing 6301 of compressor 6300.

FIG. 40 Side view of compressor housing 6301 taken along plane A-A ofFIG. 39.

FIG. 41 (a) Top view of top housing end plate 6303 of compressor 6300;and (b) side view of top housing end plate 6303 taken along plane A-A of(a).

FIG. 42 (a) Top view of bottom housing end plate 6305 of compressor6300; and (b) side view of bottom housing end plate 6305 taken alongplane B-B of (a).

FIG. 43 (a-j) Schematic top views of actuated flap compressor 7300 indifferent stages of its rotational cycle.

FIG. 44 Top view of inner drum 7302 and vane connecting rod 7370 ofcompressor 7300.

FIG. 45 Side view of shaft configuration to permit two axes of rotationof compressor 7300.

FIG. 46 Perspective view of rod shroud 7390 of compressor 7300.

FIG. 47 (a-p) Schematic views of orbiting scrolls of compressor 8000 indifferent stages of the orbiting cycle.

FIG. 48 Schematic cross sectional view of scroll compressor 8000incorporated into cooler 8800.

FIG. 49 Schematic cross sectional view of scroll compressor 8400incorporated into cooler 8801.

FIG. 50 Schematic perspective view of a gear arrangement for creatingorbiting motion.

FIG. 51 Cutaway side view of the gear arrangement of FIG. 50.

FIG. 52 Schematic cross sectional view of compressor 8500 incorporatedinto cooler 8802.

FIG. 53 (a-j) Schematic top views of gerotor compressor 9300 indifferent stages of its rotational cycle.

FIG. 54 Top view of top inlet end plate 9303 of compressor 9300.

FIG. 55 Top view of bottom outlet end plate 9305 of compressor 9300.

FIG. 56 Schematic cross sectional view of gerotor compressor 9300 (withgear box).

FIG. 57 Top view of outer gerotor 9308 and plate 9320.

FIG. 58 Perspective view of variable port mechanism 9359.

FIG. 59 Perspective view of variable port mechanism 9369.

FIG. 60 Perspective view of variable port mechanism 9379.

FIG. 61 Side view of another embodiment of a variable dischargemechanism.

FIG. 62 Top view of the variable discharge mechanism of FIG. 61 beingdriven by a servo motor.

FIG. 63 Schematic cross sectional view of gerotor compressor 9400.

FIG. 64 Top view of gears 9461 and 9460 and gerotors 9402 and 9408.

FIG. 65 Schematic cross sectional view of gerotor compressor 9500.

FIG. 66 Top view of outer gerotor 9508 and coupling plate 9320.

FIG. 67 Top view of inner gerotor 9502.

FIG. 68 Schematic cross sectional view of gerotor compressor 10300.

FIG. 69 Cross sectional view of vacuum pump 10060.

FIG. 70 Perspective view of piston 10610 of vacuum pump 10060.

FIG. 71 Schematic cross sectional view of gerotor compressor 10300incorporated into cooler 10000.

FIG. 72 Schematic cross sectional view of gerotor compressor 10400.

FIG. 73 Schematic cross sectional view of gerotor compressor 11400incorporated into cooler 11000.

FIG. 74 (a) Schematic cross sectional view of pivotable mount 11490; (b)top perspective view of top housing plate 11403 for use with pivotablemount 11490; (c) top perspective view of rotating plate 11480 for usewith pivotable mount 11490; and (d) top perspective view of ring 11491,hub 11492 and spokes 11493 of mount 11490.

FIG. 75 Cross sectional schematic view of a variable discharge portoperated by a bellows.

FIG. 76 Top view of the port of FIG. 75.

FIG. 77 Cross sectional view of a discharge mechanism control employinga liquid-containing bulb.

FIG. 78 Depiction of the removal of noncondensibles using packing andchilled water.

FIG. 79 Schematic cross sectional view of compressor 11400 and vacuumpump 12060 incorporated into cooler 12000.

FIG. 80 (a) Top view of gerotor vacuum pump 12060; and (b) side view ofgerotor vacuum pump 12060.

FIG. 81 Schematic of vacuum pump 12402.

FIG. 82 Schematic of vacuum pump 12403.

FIG. 83 Schematic of a multistage vapor compression evaporative cooler13000.

FIG. 84 Energy analysis of a multistage evaporative cooler withoutturbines.

FIG. 85 Energy analysis of multistage evaporative cooler with turbines.

FIG. 86 Generalized compressor chart.

FIG. 87 Table depicting results of a centrifugal compressor analysis.

FIG. 88 Schematic of vapor compression evaporative cooler 13100 whichemploys multistage evaporators.

FIG. 89 Schematic of vapor compression evaporative cooler 13200 whichemploys multistage evaporators and condensers.

DESCRIPTION OF THE INVENTION

This invention is directed to highly efficient and economical vaporcompression evaporative coolers that use water rather than CFCs as acoolant. Such coolers can assume various configurations such as thefollowing novel cooling systems described herein:

1. The two cooling systems depicted in FIGS. 1 and 2 which use acompressor to pressurize water vapors, followed by a jet condenser orspray condenser;

2. Cooling systems having three concentric chambers such as cooler 200depicted in FIG. 5 and cooler 3200 depicted in FIG. 16;

3. Cooling systems such as coolers 8800, 8801 and 8802 depicted in FIGS.48, 49, and 52 with condenser on top and compressor discharge port atthe top;

4. Cooling systems such as coolers 10000, 11000, and 12000 depicted inFIGS. 71,73, and 79 with condenser on bottom and compressor dischargeport on bottom; and

5. Multistage systems, such as the three systems 13000, 13100, and 13200depicted in FIGS. 83, 88 and 89.

The invention is also directed to novel low-friction, positivedisplacement compressors that can be incorporated into one or more ofthe foregoing cooling systems, including systems having three concentricchambers, such as cooler 3200 or cooler 200. These compressors include:

1. The piston compressor 230 depicted in FIG. 5;

2. Sliding vane compressor 3300 depicted in FIGS. 6-16;

3. Sliding vane compressor 4300 depicted in FIGS. 17-24;

4. Actuated sliding vane compressor 5300 depicted in FIGS. 25-31;

5. Actuated flap compressor 6300 depicted in FIGS. 32-42; and

6. Activated flap compressor, multi-vane configuration 7300, depicted inFIGS. 43-46.

This invention is also directed to the use of novel low-friction,positive displacement compressors in one or more of the above-describedcooling systems, including systems having two concentric chambers, suchas cooler 8800 depicted in FIG. 48. These compressors include:

1. Scroll compressors such as the three embodiments depicted in FIGS.48, 49, and 52 (8000, 8400, and 8500); and

2. Gerotor compressors such as the three embodiments depicted in FIGS.56, 63, and 65 (9300, 9400, and 9500).

This invention is also directed to the use of novel low-friction,positive displacement compressors that can be used in one or more of theabove-described cooling systems, such as coolers 10000, 11000, and12000. These compressors include gerotor compressors having novel meansfor actuating said compressors such as:

1. Gerotor compressor 10300, depicted in FIGS. 68 and 71;

2. Gerotor compressor 10400, depicted in FIG. 72; and

3. Gerotor compressor 11400, depicted in FIG. 73.

The present invention is also directed to novel means for removingnoncondensibles from vapor compression evaporative coolers. These meansinclude the use of vacuum pumps such as:

1. Vacuum pump 10060 depicted in FIGS. 69 and 70, which removesnoncondensibles from the condenser;

2. Gerotor vacuum pump 12060 depicted in FIG. 80 a;

3. The noncondensible vacuum pumps 12402 and 12403 depicted in FIGS. 81and 82; and

4. Pump 8060 depicted in FIG. 48.

The embodiments of the present invention are illustrated in FIGS. 1-89,wherein like numerals are used to denote like elements.

FIG. 1 depicts vapor-compression evaporative cooler 100 in accordancewith a first embodiment of the present invention. This cooler can beused in any region of the country regardless of the humidity; however,its performance is enhanced in regions with particularly low humidity.

As depicted in FIG. 1, room air (about 25° C. dry-bulb temperature, 55%relative humidity, 15° C. dew point) enters room air contactor 102through room air inlet 103. Cold water 104 (about 13° C.) is sprayedinside room air contactor 102. Room air contactor 102 has a room aircontactor return 105. The room air becomes chilled due to the directcontact with cold water spray 104. Further, the room air is alsodehumidified because the cold water temperature is less than the air dewpoint. FIG. 1 depicts room air contacting water in a spray tower 106;however, contact could also be accomplished by blowing room air througha packed tower. In a preferred embodiment, structured packing isemployed consisting of corrugated chlorinated polyvinyl chloride (CPVC)sheets that are arranged with open channels allowing chilled water toflow down and room air to flow up. CPVC is a preferred material becauseit is inexpensive, lightweight, and resists degradation by ozone, whichmay be introduced to kill potential microorganisms. Alternatively, afibrous mat may be employed with cold water trickling over the fibers,or using any other suitable means, such as random packing made ofceramic, metal, or plastic.

Evaporator 120 in FIG. 1 is preferably held at a low pressure(preferably about 0.015 atm) using compressor 130, or any positivedisplacement compressor. Water from the room air contactor 102 is drawninto the evaporator 120 where it flashes and cools. This cold water ispumped out of evaporator 120 into room air contactor 102 using pump 110.

Compressor 130 pressurizes the water vapors and discharges them to acondenser, such as jet condenser 140. When compressed, the water vaporssuperheat which increases the work requirement. This can be overcome byspraying liquid water 131 directly into compressor 130 as described inU.S. Pat. No. 5,097,677 assigned to Texas A&M University, andincorporated herein by reference. Because it takes time for the water toevaporate and cool the vapors, it is best to perform the compression ina series of small steps, or to operate the compressor slowly, or toemploy very fine water droplets. The compression is preferably doneusing a low-friction positive displacement compressor (piston, vane,flap, scroll, gerotor) such as those disclosed herein, or any suitablemeans. Because of the large volume of water vapors that must becompressed, the compressor is necessarily large. To achieve highefficiency, it is essential that the compressor have low friction.

The compressed water vapors exiting compressor 130 are sent to jetcondenser 140. Jet condenser 140 operates like a venturi. High-pressureliquid water enters jet condenser 140. The throat 141 of jet condenser140 narrows causing the water velocity to increase. The kinetic energyneeded to accelerate the water comes at the expense of pressure energy,so a vacuum is produced. The high-velocity, low-pressure water is colderthan the water vapors exiting compressor 130. When these vapors contactthe high-velocity water stream, they condense onto the water stream andare swept out with the water. The diameter of lower throat 142 of thepipe exiting jet condenser 140 increases so the water velocitydecreases. This converts the kinetic energy back into pressure energy sothe water can exit at atmospheric pressure. Any noncondensibles are alsoswept out of the system.

The liquid exiting jet condenser 140 is sent to ambient air contactor150. Ambient air contactor 150 has an ambient air inlet 152 and anambient air return 153. In a preferred embodiment, it has a salt waterpurge 158. As water evaporates into the ambient air, it cools toapproach the wet-bulb temperature. Because the wet-bulb temperature isusually significantly less than the dry-bulb temperature, heat isrejected at a much lower temperature than with conventionalvapor-compression air conditioners. In addition, because direct contactheat exchange is employed, the ΔT is much less. Cooled water is returnedfrom the ambient air contactor to the condenser via pump 112.

Because heat is ultimately rejected by water evaporation, make-up wateris required. Make-up water 131, 145 and 125 may be added to compressor130, jet condenser 140, and evaporator 120, respectively, as needed.

Some water will condense out of the room air, but this is insufficientto meet the total water requirement. If ordinary city water is employed,salts will accumulate, therefore, salt water purge 108 is used. Asindicated in FIG. 1, salt water purge 108 may be located at the bottomof the room air contactor 102. In addition, means for removingmicroorganisms from the water in the system, particularly the room aircontactor may be used, such as an ozone generator, UV radiation source,antimicrobial chemicals or other means known in the art.

FIG. 2 depicts an alternative embodiment of the present invention.Reference numerals in this embodiment refer to like elements or featuresin the embodiment of FIG. 1, so that a further description thereof isomitted. Vapor compression evaporative cooler 101 is identical to thecooler in FIG. 1, except that spray condenser 160 is employed ratherthan a jet condenser. Make-up water 165 may be added to the spraycondenser. Additionally, there is a small aspirator 170, which operatesas a venturi; the reduced pressure at the venturi throat drawsnoncondensibles out of the condenser.

The embodiment shown in FIG. 2 is easier to analyze than that in FIG. 1because actual performance data is needed for the jet condenser.Therefore, the following analysis is for FIG. 2.

The coefficient of performance, COP, is defined as the heat removed inthe evaporator divided by the total work required to operate the system$\begin{matrix}{{COP} = \frac{Q_{evap}}{W_{comp} + W_{P\quad 1} + W_{P\quad 2}}} & (1)\end{matrix}$where

Q_(evap)=heat removed in the evaporator

W_(comp)=compressor work

W_(P1)=work of pump 1

W_(P2)=work of pump 2.This expression can be inverted as $\begin{matrix}\begin{matrix}{\frac{1}{COP} = {\frac{W_{comp}}{Q_{evap}} + \frac{W_{P\quad 1}}{Q_{evap}} + \frac{W_{P\quad 2}}{Q_{evap}}}} \\{= {\frac{1}{{COP}_{comp}} + \frac{1}{{COP}_{P\quad 1}} + \frac{1}{{COP}_{P\quad 2}}}}\end{matrix} & (2)\end{matrix}$The compressor COP_(comp) is $\begin{matrix}{{COP}_{comp} = {\frac{Q_{evap}}{W_{comp}} = {\eta_{ref}\eta_{compressor}\eta_{motor}{COP}_{C}}}} & (3)\end{matrix}$where

η_(ref)=refrigeration efficiency relative to Carnot efficiency (FIG. 21,Reducing Energy Costs in Vapor-Compression Refrigeration and AirConditioning Using Liquid Recycle—Part II. Performance, Mark Holtzapple,ASHRAE Transactions, Vol. 95, Part 1, 187-198 (1989) $\begin{matrix}{{\eta_{compressor} = {{compressor}{\quad\quad}{efficiency}\quad( {60 - {70\%\quad{according}{\quad\quad}{to}\quad{{Fig}.\quad 86}}} )}}{\eta_{motor} = {{motor}\quad{efficiency}\quad( {{80\%},{{although}{\quad\quad}{higher}\quad{is}{\quad\quad}{possible}}} )}}{{COP}_{C} = {\frac{E_{E\quad 1}}{T_{C\quad 2} - T_{E\quad 1}} = {{Carnot}\quad{coefficient}{\quad\quad}{of}\quad{performance}}}}} & (4)\end{matrix}$

T_(E1)=water temperature exiting evaporator (absolute temperature)

T_(C2)=water temperature exiting spray condenser (absolute temperature)

The pump coefficient of performance, COP_(P1) is given by$\begin{matrix}{{COP}_{P\quad 1} = \frac{Q_{evap}}{W_{P\quad 1}}} & (5)\end{matrix}$The pump work is $\begin{matrix}{W_{P\quad 1} = \frac{m_{1}\quad\Delta\quad P_{1}}{\eta_{pump}\rho}} & (6)\end{matrix}$where

m₁=mass flow of water through pump 1

ΔP₁=pressure increase from pump 1

ρ=water density

η_(pump)=pump efficiency (assumed to be 50% with motor losses included)

The mass flow of water is determined by performing an energy balance$\begin{matrix}{{m_{1}{C_{p}( {T_{E\quad 2} - T_{E\quad 1}} )}} = Q_{evap}} & (7) \\{m_{1} = \frac{Q_{evap}}{C_{p}( {T_{E\quad 2} - T_{E\quad 1}} )}} & (8)\end{matrix}$where

C_(P)=heat capacity of liquid water

T_(E2)=water temperature exiting room air contactorEquation 8 can be substituted into Equation 6 which in turn issubstituted into Equation 5 to give the pump coefficient of performance$\begin{matrix}{{COP}_{P\quad 1} = \frac{\eta_{pump}{C_{p}( {T_{E\quad 2} - T_{E\quad 1}} )}\rho}{\Delta\quad P_{1}}} & (9)\end{matrix}$A similar expression can be derived for the coefficient of performancefor pump 2 $\begin{matrix}{{COP}_{P\quad 2} = \frac{\eta_{pump}{C_{p}( {T_{C\quad 2} - T_{C\quad 1}} )}\rho}{\Delta\quad P_{2}}} & (10)\end{matrix}$where

T_(C1)=water temperature exiting ambient air contactorEquations 3, 9 and 10 may be substituted into Equation 2 to determinethe coefficient of performance of the entire system $\begin{matrix}{{COP} = \lbrack {\frac{T_{C\quad 2} - T_{E\quad 1}}{\begin{matrix}{\eta_{ref}\eta_{compressor}} \\{\eta_{motor}T_{E\quad 1}}\end{matrix}} + \frac{\Delta\quad P_{1}}{\begin{matrix}{\eta_{pump}C_{p}} \\{( {T_{E\quad 2} - T_{E\quad 1}} )\rho}\end{matrix}} + \frac{\Delta\quad P_{2}}{\begin{matrix}{\eta_{pump}C_{p}} \\{( {T_{C\quad 2} - T_{C\quad 1}} )\rho}\end{matrix}}} \rbrack^{- 1}} & (11)\end{matrix}$

The room air is assumed to have the following properties:

dry bulb temperature=25° C.

relative humidity=55%

dew point=15° C.

The following parameters were used to determine the COP according toEquation 11:

-   -   η_(ref)=0.97 (from FIG. 21, Reducing Energy Costs in        Vapor-Compression Refrigeration and Air Conditioning Using        Liquid Recycle—Part II. Performance, Mark Holtzapple, ASHRAE        Transactions, Vol. 95, Part 1, 187-198 (1989)    -   η_(comp)=0.7 (from FIG. 86)    -   η_(motor)=0.8    -   η_(pump)=0.5    -   C_(p)=4188 J/(kg·° C.)    -   ρ=1000 kg/m³    -   ΔP₁=1 bar=10⁵ N/m²    -   ΔP₂=1 bar=10⁵ N/m²    -   T_(E1)=13° C.=286.15 K    -   T_(E2)−T_(E1)=ΔT (for simplicity)    -   T_(C2)−T_(C1)=ΔT (for simplicity)    -   T_(C2)=T_(C1)+ΔT        FIG. 3 shows the COP under a variety of environmental        conditions. The X axis is the ambient wet-bulb temperature (°        C.). The Y axis is the coefficient of performance calculated        using Equation 11. The ΔT that results in the greatest system        efficiency is 4° C.

The coefficient of performance for a conventional R-12 air conditioningsystem is $\begin{matrix}{{COP} = {\eta_{ref}\eta_{comp}\eta_{motor}\frac{T_{E}}{T_{C} - T_{E}}}} & (12)\end{matrix}$where η_(comp) is the compressor efficiency (assumed to be 0.7),η_(motor) is the motor efficiency (assumed to be 0.8), T_(E) is theevaporator temperature, T_(C) is the condenser temperature, and η_(ref)is given by FIG. 2 in the paper Reducing Energy Costs inVapor-compression Refrigeration and Air Conditioning Using LiquidRecycle—Part I. Comparison of Ammonia and R-12, Mark Holtzapple ASHRAETransactions, Vol. 95, Part 1, 179-186 (1989).

The evaporator temperature is assumed to be 10° C., which is 5° C.cooler than the room air dew point and 15° C. cooler than the room airdry-bulb temperature. FIG. 4 shows the COP for R-12 vapor compressionrefrigeration using a variety of dry-bulb temperatures and condenser ΔT.The X axis is the ambient dry-bulb temperature (° C.). The Y axis is thecoefficient of performance calculated by Equation 12.

Table 1 compares the expected performance of the vapor-compressionevaporative cooler and the conventional R-12 vapor compression system ina variety of cities in the United States. The dry bulb and wet bulbtemperatures are the “2.5 values,” meaning only 2.5% of the hoursbetween June and September exceed these values. From this analysis, itis expected that the vapor-compression evaporative cooler is 1.7 to 3.9times more energy efficient than conventional vapor-compression airconditioning. This energy comparison does not include energy for theambient air fan or the room air blower. TABLE 1 Comparison ofConventional Air Conditioning to Vapor-Compression Evaporative CoolerVapor- Compression R-12 Vapor Evaporative Compression† Cooler City DryBulb Temp (° C.) COP₁ Wet Bulb Temp (° C.) COP₂$\frac{{COP}_{2}}{{COP}_{1}}$ Albuquerque, 33 4.4 16 15.0 3.4 New MexicoAtlanta, Georgia 33 4.4 23 8.8 2.0 Dallas, Texas 36 3.9 24 8.2 2.1 ElPaso, Texas 37 3.8 18 12.4 3.3 Houston, Texas 34 4.2 25 7.8 1.9 LasVegas, Nevada 41 3.2 18 12.4 3.9 Los Angeles, California 32 4.5 21 10.02.2 Miami, Florida 32 4.5 25 7.8 1.7 Minneapolis, Minnesota 37 3.8 238.6 2.3 New Orleans, Louisiana 33 4.4 26 7.3 1.7 New York, New York 324.5 23 8.8 2.0 Phoenix, Arizona 42 3.2 22 9.4 2.9 Sacramento, California37 3.8 21 10.0 2.6 Salt Lake City, Utah 35 4.1 17 13.2 3.2 Spokane,Washington 32 4.6 17 13.2 2.9 Washington, D.C. 33 4.4 23 8.6 2.0†Calculations assume the condenser temperature difference is 10° C.

Another embodiment of the invention is directed to vapor-compressionevaporative coolers having three concentric chambers. One such cooler iscooler 200 depicted in FIG. 5. The innermost chamber 210 ofvapor-compression evaporative cooler 200 is at the lowest pressure, themiddle chamber 211 is slightly higher, and the outermost chamber 212 isat atmospheric pressure. The outer diameter of the outermost chamber ispreferably two to three feet for a 3-ton home air conditioning unit andpreferably stands about three to four feet high. At the top of theoutermost and middle chambers and below compressors 230 in the innermostchamber 210 are circular pipes 214 (inner), 216 (middle), and 218(outer) through which water drips. If desired, packing 220 and 221 canbe placed in the middle and outer chambers to increase the water holdup.

One or more compressors 230 are preferably disposed in innermost chamber210 creating a vacuum in the chamber. As a result, water in theinnermost chamber 210 evaporates and becomes chilled. This chilled water224 is pumped into room air contactor 1000 located in the house or thespace to be cooled where it countercurrently contacts warm air such thatthe air then becomes cooled. The chilled water is sufficiently low intemperature that humidity in the house air will actually condense; thus,not only is the house air cooled, but it is also dehumidified. A furtherbenefit is that the house air is literally washed which removes dust andallergens.

Because water is evaporating in the innermost chamber, it must bereplaced. This is preferably accomplished by float 226 which opens avalve 227 allowing tap water to be introduced to replenish theevaporated water. Because tap water contains salts, a purge stream 228is be used to remove salt as it becomes concentrated. This may beaccomplished by opening valve 229 when the salt concentration exceeds agiven level. The valve can be opened based on a timer, a waterconductivity meter, purging a continuous flow rate known to be suitablefor the tap water salt concentration, or other suitable means known inthe art. Alternatively, distilled water or rain water could be used asthe system water and as make-up water such that purging would becomeunnecessary. However, in such a system, the water must be completelysalt free.

The pressure in the innermost chamber is kept low using one or morecompressors 230. Although FIG. 5 depicts two reciprocating compressorsoperating in parallel, it should be understood that any suitablecompressor may be used. Particularly suitable are low-friction positivedisplacement compressors such as the disclosed reciprocating compressor(FIG. 5), sliding vane compressors (FIGS. 6-14, 16 and 25-31) andactuated flap compressors (FIGS. 32-42 and 43-46).

In the embodiment depicted in FIG. 5, reciprocating compressor 230 isused. Because the vapor density is very low, the compressor feature mayconsist of many stages. For example, in FIG. 5, two are shown operatedin parallel. Alternatively, multiple compressors may be operated inseries, as shown in FIGS. 83, 88, and 89. In FIG. 5, the compressorcenter shaft 232 oscillates vertically. The top of the shaft has aprotrusion 234 which rides in a sinusoidal groove 237 in rotating cam236. A single sinusoid in groove 237 causes the center shaft to cycleonce per motor rotation. By placing a groove with two sinusoids in cam236, the center shaft will oscillate twice per motor rotation. Thus,very rapid center-shaft oscillations are possible using a conventionallow-speed motor 238.

The top end of bellows 240 is affixed to the oscillating shaft 232 andthe other end is affixed to housing 242, thus maintaining a vacuum-tightseal. The center shaft 232 has plates 244 attached to it that alsooscillate vertically within compressor housing 290.

It has been discovered that because of the lower pressures involved inthe cooling systems of the present invention, it is not necessary forclose contact between the compressor components such as the plates 244and housing 290. Thus, in the present embodiment, plates 244 preferablyhave a sizable gap 245 (a few thousandths of an inch) with the housing290 so they do not touch the housing giving negligible friction.Instead, water is used to make a seal. Specifically, the edges of theplates contain a groove 246 through which tap water flows. Because thetap water is at a higher pressure than the compressor, liquid waterflows into the compressor housing 290 rather than vapors leaking out.This water not only provides a seal, but it also cools the compressedvapors. If this water is insufficient to cool the compressed vapors,additional water spray nozzles 248 may be placed on the face of theplates 244. The source of the tap water is through the center shaft 232which is hollow and has a flexible hose 233 at the bottom. Thecompressor has an outlet 249 to the middle chamber, regulated by outletvalve 250. The compressor also has an inlet 251 regulated by inlet valve252. Because the pressures are so low, it is not feasible to open valves250 and 252 using pressure differences between the compressor interiorand exterior. Instead, the inlet valve 252 and outlet valve 250 areactive, i.e., actuated by electric solenoids or hydraulic pistons. Thecompressor housing 290 preferably has a slightly conical bottom 243 sothat excess liquid collects at the outlet valve 250 to be dischargedwhen the valve opens. The outlet 249 of the compressor is directed tothe middle chamber 211.

In the middle chamber 211, the vapors condense directly onto the waterspray 217 from circular pipe 216. If too much liquid collects at thebottom of the middle chamber, an electrical switch 256 is activated by afloat 254 which turns on pump 258 to remove liquid. If the water leveldrops too low, the float 254 turns the switch 256 off. Becausenoncondensible gases will collect in the middle chamber 211, they may bepurged by aspirator 270.

The liquid that is pumped out of the middle chamber 211 goes to theouter chamber 212 where it contacts ambient air and is cooled toapproach the wet-bulb temperature. The air is blown by fan 272 locatedat the top of the unit. The fan 272 and compressor 230 are preferablyboth powered by the same motor 238 which reduces costs and increasesefficiency. Further, the fan 272 acts as a flywheel. Float 274 operatesa valve 275 which introduces make-up water to the outer chamber 212 asrequired. To purge concentrated salts, valve 276 opens periodically topurge some of the liquid out of the outer chamber 212. An ozonegenerator or other means known in the art may be used to inhibitmicroorganism growth in the system.

In a preferred embodiment, the following parameters may be used:

1. motor speed=1725 rpm

2. cam causes one center-shaft oscillation per motor rotation

3. the compressor has an 80% volumetric efficiency

4. three stages operating in parallel

5. 3-in stroke

6. a 3-ton unit must compress 1400 ft³/min of low-pressure vapors.

In the embodiment depicted in FIG. 5, the diameter of plate 244 would be1.3 ft. Smaller diameters are possible by using a higher speed motor,altering the cam to allow more center-shaft oscillations per motorrotation, increasing the stroke, or increasing the number of stages.

Regulation of the system may be accomplished by on-off control as isdone with conventional air conditioners. Alternatively, a variable-speedmotor may be used to operate the compressor more efficiently; when theload is low, it runs more slowly and when the load is high, it runs morequickly.

Because the pressure differences across all walls are very low (15 psiat most), almost all of the components may be constructed of plasticthus reducing cost. However, any suitable material may be used to makethe individual components. Because none of the chambers is completelyfilled with liquid water, if the water were to freeze in winter, the 10%volume increase of the ice can be accommodated by the vapor space. Ifdesired, the unit could be drained of water to winterize it.

This embodiment may also be adapted to below-freezing applications byadding a nonvolatile antifreeze (e.g., salt, glycerol) to the water.This will lower the vapor pressure of the water thus requiring a largercompressor. In addition, if antifreeze is added to the water, then allmake-up water should be distilled water so that no salts must be purged.Alternatively, the complexity and cost associated with adding distilledwater may be eliminated if tap-water salts are used as the antifreeze.This could be accomplished simply by operating the system with a verylow purge rate.

The present invention is also directed to other types of positivedisplacement compressors that may be used in vapor-compressionevaporative coolers, including coolers having three concentric chambers.Rotary compressors are one type of compressor which can be used in suchcoolers. One embodiment of a rotary compressor useful in vaporcompression evaporative coolers, rotary sliding vane compressor 3300, isdepicted in FIGS. 6-14. In FIGS. 6-14, like reference numerals refer tolike elements. The novel rotary compressors of this and later disclosedembodiments employ a sizable gap to reduce friction between thecompressing components and use water both as a sealant and a coolant.

FIGS. 6 a-f are a schematic of rotary sliding vane compressor 3300 invarious stages of the cycle. As the rotor 3302 rotates, it sweeps invapors from the low-pressure side 3314 during the first rotation andthen compresses them during the second rotation. To cool the vaporsduring the compression and to make seals, liquid water 3306 is sprayedinto the compressor during the compression. As shown in FIGS. 6 a-f,sliding vane 3308 contacts the rotor 3302 and separates the low-pressure3314 and high-pressure 3315 sides of the compressor. Alternatively, toreduce friction, the sliding vane 3308 could contact the rotor 3302through a roller located at the tip of the sliding vane 3308, or thesliding vane 3308 could be actuated by an external mechanism so therotor 3302 and sliding vane do not touch.

FIG. 7 is a three dimensional schematic of sliding vane rotarycompressor 3300. The low-pressure vapors enter inlet hole or port 3310in the side of compressor housing 3301. No inlet check valve isrequired. The high-pressure vapors exit through outlet or discharge port3312. Outlet check valves 3313 (FIG. 13) are provided at the outlet.

FIG. 8 is a side view of rotor 3302. As depicted in FIG. 8, rotor 3302consists of a top cap 3320, bottom cap 3330 and sidewall 3340 which ispreferably cylindrical. Top sealing groove 3322 and bottom sealinggroove 3332 located on the top and bottom caps are filled with liquidwater to create a rotating seal against the housing end plates 3303(FIG. 14 a) and 3305 (FIG. 15A). A vertical groove 3342 on the sidewallseals the rotor 3302 against the compressor housing 3301. The groovesmay be entirely open, or they may contain a wick 3324 along the open endas depicted in FIG. 9, a drawing of the groove detail with a wick.Grooves 3322, 3332 and 3342 may be actively supplied with low-pressurewater 3327 through a slip ring 3326. Distribution channels 3329 ensurethat water is distributed to sealing grooves. Alternatively, the watersource may be from the water spray 3306 used to cool the compressor. Thewater spray will wet the interior walls of the compressor housing and bewicked into the grooves 3322, 3332, and 3342, provided wick 3324 isused.

Because the rotor is eccentrically mounted on the shaft, a counterweight3328 is needed to balance the rotation. FIG. 8 shows the counterweight3328 may be located inside the sidewall 3340 to save space.Alternatively, the counterbalance or counterweight 3328 may be locatedon the shaft outside the compressor housing.

FIGS. 8-10 show details of the components of rotor 3302. Top and bottomend caps 3320 and 3330 each have a large hole 3321 (top) and 3331(bottom) to reduce the mass that must be counterbalanced. The largeholes also provide a means to drain the water from the rotor interior.

FIGS. 11 a-c show the details of sliding vane 3351. It has pins 3352located on the interior which ride on linear bearings 3354. The slidingvane 3351 is forced against the rotor using springs 3356. Felt wiper3358 becomes water soaked from the cooling water spray 3306, so it sealsagainst the rotor. A novel feature of this embodiment is the means usedto create the gap between the felt wiper 3358 and the rotor.Specifically, roller bearings 3360, which protrude farther than the feltwiper, ride on the rotor. The clearance between the felt wiper and therotor is determined by the amount the roller bearings 3360 protrude fromthe wiper 3358. As can be seen from the foregoing, there is low frictionbetween the compressing components of the disclosed embodiment—therotor, housing and sliding vane.

FIGS. 12-13 show the compressor housing 3301. It contains a section 3361for sliding vane 3351; sealing grooves 3362 on the side provide a seal.Inlet port 3310 is entirely open, but discharge port 3312 is lined withcheck valves 3313. Because the pressure generated by the compressor isnot sufficient to actuate the check valves, they are preferablyactivated by solenoids, hydraulic pistons, or other means. Excess waterwill collect in the water sump 3363 which discharges through sump checkvalve 3364.

The housing end plates 3303 and 3305 are depicted in FIGS. 14-15. Topend plate 3303 and bottom end plate 3305 have top bearing cup 3307 andbottom bearing cup 3309. Bottom end plate 3305 also has a drainage hole3311.

FIG. 16 depicts sliding vane rotary compressor 3300 integrated intovapor-compression evaporative cooler 3200. Like cooler 200, it has threeconcentric chambers. Drive motor 3238 may be located inside or outsideof the evaporator. In a preferred embodiment, it is outside. Locatingthe drive motor outside the evaporator has the following advantages: 1)waste heat will not put a load on the compressor, 2) a standard motorcan be employed rather than one specially designed for use in alow-pressure, water-vapor environment, and 3) easy servicing. Drivemotor 3238 is coupled to rotor 3302 by rotary shaft 3232.

A rotary shaft seal 3233 is required. The present invention is alsodirected to a novel rotary shaft seal useful in compressor 3300 as wellas other applications. Specifically, as depicted in FIG. 16, a seal isprovided by supplying journal 3240 with water. Water will be drawn intothe evaporator 3341 because it is at a low pressure. Provided excesswater is supplied to journal 3240, no air will leak into the evaporator3341. The clearance between shaft 3232 and journal 3240 may berelatively large so there is low friction. To prevent air from leakinginto evaporator 3341 when motor 3238 is off, a bellows seal 3339 isemployed. Due to centrifugal force, the bellows seal 3339 lifts off thejournal face 3241 when shaft 3232 rotates, but seats on the journal face3241 once the shaft rotation stops. Using this arrangement, there isvery little friction due to the shaft seal. Although seal 3233 isdescribed in connection with compressor 3300, as will be clear to thoseof skill in the art, it may be also used in other applications.

The sliding vane compressor 3300 pulls a vacuum on the evaporator 3341causing the liquid water to evaporate. Nucleation sites (e.g., “boilingchips”) 3337 are preferably provided to increase the evaporationefficiency. As water 3224 evaporates, it becomes chilled. This chilledwater is pumped out of the evaporator 3341 and into a room air contactor3102 via pump 3502. House air directly contacts the chilled water whichcools it and removes humidity.

The vapors discharged from the compressor 3300 enter the condenser 3211which has water trickling over condenser packing 3220. The inlet wateris near the wet-bulb temperature of the ambient air, which is coolerthan the compressor discharge temperature, so the vapors condense ontothe packing 3220. The packing may be a structured packing consisting ofcorrugated plastic or metal sheet, or a random packing such as ceramicsaddles. A pump 3503 removes the warmed water from the condenser anddirects it to the ambient air contactor 3212. Moisture evaporates fromthe cooling of the water so it can be reintroduced into the condenser3211. To facilitate contact between the ambient air and the warm water,the ambient air contactor 3212 may have a structured or random packing3214.

Tap water which may be used to cool the compressor 3300 passes through aheat exchanger 3221 in contact with the water in the ambient aircontactor. This step is necessary only if the tap water temperature isgenerally above the wet-bulb temperature.

As an option, a pump 3500 can be added which pumps water out of theambient air contactor 3212 and sends it to the home refrigeratorcondensing coil (not shown.) This will increase the refrigeratorefficiency because: 1) water has better heat transfer properties thanair, and 2) the water temperature will generally be lower than roomtemperature. Water returning from the refrigerator may be directed backto the ambient air contactor. Pump 3400 is used to flow water throughaspirator 3270 in order to remove noncondensibles from condenser 3211.

To regulate the water levels in the various tanks, float valves 3227(inner), 3275 a (outer), 3275 b (outer) and 3256 (middle) may beemployed. Most of the float valves introduce water into the tank if thewater level drops too low. An exception is the left float valve 3275 ain ambient air contactor 3212. Because water is constantly being addedto the ambient air contactor, it will tend to fill up. The left floatvalve 3275 a is designed to open when the water level gets too highallowing water to be sucked into condenser 3211. The right float valve3275 b in the ambient air contactor 3212 is necessary only if water issent to the refrigerator. During the winter, the various water make-upsystems would not be used because the air conditioner is not required.However, due to the refrigeration load, water will evaporate from theambient air contactor thus dropping the liquid level. When the waterlevel drops, the right float 3274 b opens a valve 3275 b allowingmake-up water to be introduced.

Because noncondensible gases will accumulate in condenser 3211, anaspirator 3270 is used to pump out the gases. The motive force foraspirator 3270 is provided by circulating pump 3400. Alternatively, amechanical vacuum pump could be employed. For instance, vacuum pumps12060, 12402, 12403 and 10060, discussed below, could be employed.

Because the evaporator 3341 and condenser 3211 are operated at very lowpressures, pumps 3502 (evaporator) and 3503 (condenser) are provided toremove liquid from these vessels. However, liquid that enters thesevessels requires no pump because they are at low pressure. Potentially,turbines may be used to capture the energy of the water as it flows intothe low-pressure vessels.

Water evaporates from both the evaporator 3341 and ambient air contactor3212, which will increase the salt concentration in the water. Water ispurged from the room air contactor 3102 and may be added to thecondenser 3211 or dumped to the sewer. Additionally, water is purgedfrom the condenser 3211 and may be sent to the sewer. The rate thatwater is purged from the system can be regulated by a pre-set valve, atimer-controlled valve, a salinity meter, or other means known in theart.

Because the evaporator is cold relative to the ambient environment,insulation 3405 is preferably used to maintain system efficiency.

Except for the oscillating mass of the sliding vane, rotary compressor3300 will be virtually vibration free. In contrast, a reciprocatingcompressor produces much vibration. In addition, reciprocatingcompressors require an inlet check valve which adds expense and lowersefficiency because of flow losses through the valve.

A centrifugal or axial compressor must operate at very high speedsrequiring either expensive high-speed motors or gear boxes. Sliding vanecompressor 3300 can operate using conventional motors. Also, high-speedcentrifugal and axial compressors may not tolerate liquid dropletsneeded to cool the compressor. A centrifugal or axial compressor will bemore expensive because it has many precision components and it must bewell balanced.

Still another embodiment of the invention is directed to anotherlow-friction sliding vane compressor useful in a vapor-compressionevaporative cooler such as cooler 3200. This compressor is depicted inFIGS. 17-24. Like sliding vane compressor 3300, this compressor useswater both as a sealant and a coolant.

FIGS. 17 a-f show a schematic of sliding vane compressor 4300 in variousstages of the cycle. As rotor 4302 rotates, it sweeps in vapors from thelow-pressure side 4314 during the first rotation and then compressesthem during the second rotation. To cool the vapors during thecompression and to make seals, liquid water 4306 is sprayed into thecompressor 4300 during compression. Sliding vane 4308 nearly contactsthe compressor housing 4301 and separates the low-pressure side 4314 andhigh-pressure side 4315 of the compressor.

FIG. 18 is a three dimensional schematic of sliding vane compressor4300. The low-pressure vapors enter inlet hole or port 4310 in the sideof the compressor housing 4301. No inlet check valve is required. Thehigh-pressure vapors exit through outlet or discharge port 4312. Outletcheck valves 4313 (FIG. 22) are provided at the outlet.

FIG. 19 is a side view of rotor 4302. As depicted in FIG. 19, rotor 4302consists of a top cap 4320, bottom cap 4330, and sidewall 4340 which ispreferably cylindrical. Top cap 4320 has a drainage hole 4323, andbottom cap 4330 has drainage hole 4333. Top sealing groove 4322 andbottom sealing groove 4332 located on the top and bottom caps are filledwith liquid water to create a rotating seal against the housing endplates 4303 and 4305 which are depicted in FIGS. 23-24. Also, as shownin FIG. 20, sidewall 4340 has a vertical gap 4366 with a groove 4368that seals against the sliding vane. Connectors 4361 are inserted intovertical slot 4366 to seal against the upper and lower surfaces ofsliding vanes 4308. The grooves 4322 and 4332 shown in FIG. 19 may beentirely open, or they may contain a wick 4324 along the open endsimilar to that depicted in FIG. 9. The grooves may be actively suppliedwith low-pressure water 4327 through a slip ring 4326. Distributionchannels 4329 ensure that water is distributed to sealing grooves.Alternatively, the water source may be from the water spray 4306 used tocool the compressor. The water spray will wet the interior walls of thecompressor housing and be wicked into grooves 4322 and 4332, providedwick 4324 is used.

The rotor components, particularly the sliding vane, have pins 4352,linear bearings 4354, springs 4356, a felt wiper 4358, and rollerbearings 4360 similar in structure and operation to those elements 3352,3354, 3356, 3358, and 3360 depicted in FIGS. 11 a-c, so a furtherdescription thereof is omitted.

FIGS. 21-22 show the compressor housing 4301. Inlet port 4310 isentirely open, but discharge port 4312 is lined with check valves 4313.Excess water will collect in the water sump 4363, which dischargesthrough sump check valve 4364. Sealing groove 4319 on housing 4301 sealsagainst the rotating sidewall 4340. The groove may be entirely open, orit may contain a wick 4324 along the open end similar to that depictedin FIG. 9.

The housing end plates 4303 and 4305 are depicted in FIGS. 23-24. Tophousing end plate 4303 and bottom housing end plate 4305 have topbearing cup 4307 and bottom bearing cup 4309. Bottom end plate 4305 alsohas drainage holes 4311.

Sliding vane compressor 4300 may be integrated into vapor-compressionevaporative cooler 3200 in place of compressor 3300 depicted in FIG. 16.Except for the oscillating mass of the sliding vane, sliding vanecompressor 4300 will be virtually vibration free. In addition, slidingvane compressor 4300 is more compact than compressor 3300.

Still another embodiment of the invention is directed to an actuatedsliding vane compressor for use in a vapor-compression evaporativecooler, such as cooler 3200. This compressor is depicted in FIGS. 25-31.Like the previous embodiments, this compressor design minimizes frictionand uses water both as a sealant and a coolant.

FIGS. 25 a-f show a schematic of actuated sliding vane compressor 5300in various stages of the cycle. As rotor 5302 spins, sliding vane 5308is positioned near the housing 5301 by roller bearings 5360 which ridein grooves 5316.

Referring now to FIGS. 25-31, actuated sliding vane compressor 5300comprises rotor 5302, which is disposed inside and rotates in compressorhousing 5301 in a similar fashion to rotor 4302 and compressor housing4301, depicted in FIGS. 17-18.

FIGS. 25-29 depict the details of the rotor components of actuatedsliding vane compressor 5300. Rotor 5302 consists of a top cap 5320, abottom cap 5330 and sidewall 5340 which is preferably cylindrical. Therotor caps have grooves 5322 and 5332. As depicted in FIG. 28, sidewall5340 has vertical grooves 5368 and vertical gap 5366, similar instructure and function to grooves 4368 and gap 4366 in the previousembodiment. Connectors 5361 are inserted into vertical slot 5366 to sealagainst the upper and lower surfaces of sliding vane 5308. Water 5327may be supplied to the grooves via slip ring 5326. Distribution channels5329 ensure that water is distributed to sealing grooves. As shown inFIG. 27, slot 5321 in top cap 5320 and slot 5331 in bottom cap 5330allow roller bearings 5360 on sliding vane 5308 to protrude from endcaps 5320 and 5330. FIGS. 29 a-c depict some of the components ofsliding vane 5308, including pins 5352 located on the interior whichride on linear bearings 5354. As depicted in FIGS. 29-31, rollerbearings 5360 of sliding vane 5308 ride in grooves 5316 located on thehousing end plates 5303 and 5305. The springs 5356 shown in FIG. 29 amay be in compression so the roller bearings 5360 of sliding vane 5308ride on the outside edge of grooves 5316. Grooves 5316 may have acircular cross-section or may be non-circular and function as a cam tocarefully regulate the position of the sliding vane. Housing end plates5303 and 5305 have bearing cups 5307 and 5309, respectively, to supportshaft 5370. Bottom housing end plate 5305 has a port 5318 to drainexcess water. A sintered metal or, alternatively, felt wiper 5358becomes water soaked from the cooling water spray, so it seals againstthe housing 5301.

The compressor housing of the present embodiment is similar in structureto housing 4301 depicted in FIGS. 21-22 of the previous embodiment, suchthat a further description thereof is omitted. As in the previousembodiment, the inlet port is entirely open, but the discharge port islined with check valves. Excess water will collect in a water sump whichdischarges through a check valve.

Actuated sliding vane compressor 5300 may be integrated intovapor-compression evaporative cooler 3200 depicted in FIG. 16 in placeof rotary compressor 3300 or 4300.

To reduce the cost associated with purchasing individual motors for eachpump, and to increase the efficiency (one large motor is more efficientthan multiple small motors), the pumps (and turbines, if used) may bemounted on the same shaft that drives the compressor 5300. The pumpsneed not have tight seals because water will leak into the evaporatorwith no major adverse consequences. The loose seals will reduce frictionand increase pump efficiency.

Actuated sliding vane compressor 5300 has an advantage over sliding vanecompressor 4300 in that the roller bearings 5360 for sliding vane 5308will actually rotate relatively slowly. In contrast, roller bearing 4360must rotate very fast, which may require the use of expensive,high-speed roller bearings.

Still another embodiment of the invention is directed to an actuatedflap compressor which may be used in a vapor-compression evaporativecooler, such as cooler 3200. This compressor is depicted in FIGS. 32-42.Like the previous embodiments, this compressor has low friction and useswater both as a sealant and a coolant.

FIGS. 32 a-f show a schematic of actuated flap compressor 6300 invarious stages of the cycle. As rotor 6302 rotates, it sweeps in vaporsfrom low-pressure side 6314 during the first rotation and thencompresses them during the second rotation. To cool vapors during thecompression and to make seals, liquid water 6306 is sprayed into thecompressor 6300 during the compression. An actuated flap 6308 nearlycontacts the interior of compressor housing 6301 and separates thelow-pressure side 6314 and high pressure side 6315 of the compressor.

FIG. 33 is a three dimensional schematic of actuated flap compressor6300. The low-pressure vapors enter inlet hole or port 6310 in the sideof the housing 6301. No inlet check valve is required. The high-pressurevapors exit through outlet or discharge port 6312. Outlet check valves6313 (FIG. 40) are provided at the outlet.

FIG. 34 is a side view of rotor 6302. As depicted in FIG. 34, rotor 6302consists of a top cap 6320, bottom cap 6330, and sidewall 6340, which ispreferably cylindrical. Slots 6321 and 6331 allow the roller bearings6360 (FIG. 36) to protrude from end caps 6320 and 6330. Top sealinggroove 6322 and bottom sealing groove 6332 located on the top and bottomcaps are filled with liquid water to create a rotating seal against thehousing end plates 6303 and 6305 (FIGS. 41 and 42). The grooves may beentirely open, or they may contain a wick along the open end asdescribed in previous embodiments. The grooves may be actively suppliedwith low-pressure water 6327 through slip ring 6326. Distributionchannels 6329 ensure that water is distributed to sealing grooves.Alternatively, the water source may be from the water spray used to coolthe compressor. The water spray will wet the interior walls of thecompressor housing and be wicked into grooves 6322 and 6332, providedwick 6324 is used.

FIGS. 35-38 show further details of the rotor components. Referring toFIGS. 35-38, actuated flap 6308 has hinge pins 6352 which fit into hingeholes 6354 and 6356 in rotor top end cap 6320 and bottom end cap 6330.Flap 6308 is forced outward by centrifugal force. As depicted in FIGS.37-38 and 41-42, roller bearings 6360 ride in top guide track 6362 oftop end plate 6303 and bottom guide track 6364 in bottom end plate 6305of the compressor housing 6301 which prevents the flap from touching thecompressor housing 6301, thus maintaining a slight gap of a fewthousandths of an inch. The outer surface 6304 of flap 6308 may becovered with cloth or felt so that water is wicked between the flap andthe compressor housing, thus forming a seal. Guideposts 6359 fit throughguide hole 6358 so that flap 6308 is actuated when roller bearings 6360mounted on axle 6366 ride in guide tracks 6362 and 6364. Rotor 6302nearly contacts housing 6301 at sealing groove 6365 which may be open,or have a wick such as that depicted in FIG. 9. As can be seen from theforegoing, the present embodiment has minimal friction between thecompressor components—the rotor, flap and housing.

FIGS. 39-42 show the compressor housing 6301. Inlet port 6310 isentirely open, but the discharge port 6312 is lined with check valves6313. Excess water will collect in the water sump 6363 which dischargesthrough sump check valve 6364. The housing end plates 6303 and 6305 areshown in FIGS. 41-42. Top end plate 6303 has a top bearing cup 6307 anda guide track 6362. Bottom end plate 6305 has a bottom bearing cup 6309and a guide track 6364.

Because the final compression pressure is not great enough to open thecheck valves, they are actively opened preferably with a solenoid orhydraulic pistons. The timing of the opening/closing may be based uponmeasurements of the evaporator and condenser temperature. A “look-up”table on a computer chip may be used to open the valves at the optimalrotation angle. The optimal rotation angle may be determinedexperimentally by varying the opening angle and measuring the onesgiving the maximum coefficient of performance under a variety ofevaporator/condenser temperatures.

As with the previously disclosed compressors 3300, 4300 and 5300,actuated flap compressor 6300 may be integrated into vapor-compressionevaporative cooler 3200 depicted in FIG. 16.

The present embodiment enjoys several advantages over other compressors.Except for the oscillating mass of the flap, compressor 6300 will bevirtually vibration free. In contrast, a reciprocating compressor willhave much vibration. In addition, a reciprocating compressor requires aninlet check valve which adds expense and lowers efficiency because offlow losses through the valve. Actuated flap compressor 6300 is alsomore compact than rotary compressor 3300.

Further, a dynamic compressor (centrifugal or axial) must operate atvery high speeds requiring either expensive high-speed motors or gearboxes. Actuated flap compressor 6300 can operate using conventionalmotors. Also, high-speed centrifugal and axial compressors may nottolerate liquid droplets needed to cool the compressor. A centrifugal oraxial compressor may also be more expensive because they have manyprecision components and must be well balanced.

Another embodiment of the invention is directed to a simple, valveless,cost-effective water vapor compressor with a variable compression ratio,which can be used in vapor-compression evaporative cooling systems, suchas cooler 3200. This low-friction compressor, depicted in FIGS. 43-46,uses multiple swinging vanes.

As depicted in FIGS. 43-46, actuated flap compressor 7300 comprises aswinging vane 7308, a rigid connecting rod 7370 for vane angle control,a multiple vane configuration which eliminates the need for a dischargevalve, and an adjustable discharge port opening 7312 for compressionratio variation.

In this embodiment, multiple vanes 7308 are actuated in a radial fashionfrom an inner drum 7302, such that the vane tips 7372 seal against anouter drum 7301 forming a cavity of decreasing volume upon rotation. Thehigh pressure side of the vane can be curved with an arc with the sameradius as the outer drum, to insure complete discharge by minimizingdead volume. Actuation means eliminates the frictional losses suffereddue to contact between the vane tips 7372 and the outer drum 7301.

FIGS. 43 a-j show the progression of an arbitrarily chosen cavitythrough the compression and discharge stages. Note that the vanes 7308do not quite touch the outer drum 7301. For clarity, the means ofachieving this actuation are not shown in this figure. The compressionoccurs by collapsing the encapsulated volume 7374 (hatched area in FIGS.43 a-j) between two successive vanes 7308 beginning immediately afterthe trailing vane passes the final intake port 7310. Compression endsand discharge begins as the leading vane passes the opening of thedischarge port 7312, allowing the compressed vapor to be expelled bycontinued volume collapse without further compression. Although only onecavity has been described, all cavities perform the same function;therefore, four of the processes described above occur per revolution.Variable compression ratio is achieved by adjusting the leading edgelocation of the discharge port 7312 circumferentially on the outer drum7301 (which determines port opening time and thus compression ratio). Asnoted, in this embodiment, the ports can be valveless.

FIG. 44 shows the configuration of the vane connecting rod 7370. As thesquare inner drum 7302 rotates about its axis 7378, a very smallclearance or gap 7380 between the vane tip 7372 and the outer drum 7301can be maintained by rotating the vane connecting rod 7370 about theouter drum axis 7382.

FIG. 45 shows the method by which the two axes of rotation can beprovided. The torque coupler 7384 is driven by a power source (such asan electric motor) and transmits the torque to the inner drum 7302. Italso provides adequate translational constraints to the offset shaft7386 which is coaxial with the outer drum 7301. All degrees of freedomare constrained in the offset shaft 7386 by proper shaft shouldering androtation constraint at the base 7388. One end of the vane connecting rod7370 is fixed to the portion of the offset shaft 7386 which is coaxialwith the outer drum 7301.

The center of the inner drum 7302 is not part of the encapsulatedcompression volume, so the penetration of the vane connecting rods 7370through the wall of the inner drum 7302 should not permit flow of thecompressed water vapor. A barrier is provided by a shroud 7390 as shownin FIG. 46. This shroud 7390 mounts on the inside of the swinging vane7308 and moves in and out of the inner drum wall 7392 as dictated by theangle between the vane 7308 and the inner drum 7302.

Actuation of the swinging vane 7308 is simple and can be achievedwithout dry friction losses caused by contact between vane 7308 and theouter drum 7301. Conventional sliding vane compressors are very small,making these frictional losses acceptable. The application of vaporcompression for an air conditioning system requires very large flowrates and thus a compressor with large geometry, making dynamic contactbetween the vanes and the outer drum unacceptably inefficient. In theswinging vane compressor, a very thin gap 7380 can be maintained betweenthe vane tips and the outer drum by a simple connecting rod as describedabove, thus eliminating contact frictional losses. Further, no checkvalves are required which greatly simplifies the design.

Still another embodiment of the present invention is directed to coolingsystems having two concentric chambers. The outer chamber contains anambient air contactor. The inner chamber is subdivided into acompression and condenser chamber with the compressor in between. In oneembodiment, the condenser chamber is disposed on the top and theevaporator chamber is on the bottom. The present invention is alsodirected to positive displacement, low-friction compressors useful insuch coolers. These include scroll compressors and rotary compressorssuch as gerotor compressors. One such scroll compressor is depicted inFIGS. 47-48. This embodiment is incorporated into a cooler that canprocess the very large volumetric flow rate of water vapor and inaddition, novel means are provided to remove noncondensibles from thesystem. In addition, the scroll compressor of the embodiment describedherein requires no valves which greatly simplifies the design.

Scroll compressor 8000 is depicted in FIGS. 47-48. FIGS. 47 a-p show asequence of images indicating changes in the volume of gas as mobilescroll 8004 orbits around stationary scroll 8003. During the first fewstages, gas is taken in at a low pressure. Once it is sealed off, thevolume reduces and the pressure rises. The high-pressure gas is releasedthrough a hole 8011 in the stationary scroll 8003.

FIG. 48 depicts a schematic cross section of scroll compressor 8000integrated into vapor-compression evaporative cooler 8800. Cooler 8800uses scroll compressor 8000 to pressurize water vapor. Electric motor8001 drives scroll compressor 8000 through flexible coupling 8002. Thescroll compressor 8000 has two stages connected in series: first stage8000 a and second stage 8000 b. As depicted in FIG. 48, the stationaryscroll 8003 of the second stage compressor 8000 b has a drive shaft 8007located on the center axis with crank 8009 which drives mobile scroll8004 in an orbital motion. Webbing 8010 provides stiffness to thestationary scroll 8003. There is a sloppy fit between the crank 8009 andthe mobile scroll 8004. Precision positioning of the mobile scroll 8004relative to the stationary scroll 8003 is obtained through rotors 8008.Although FIG. 48 shows two rotors 8008 per scroll, preferably threewould be employed. The three rotors 8008 confine the mobile scroll 8004to an orbiting motion. The rotors can be counterbalanced such that thereis no vibration in the orbiting scrolls.

Referring again to FIG. 48, first-stage compressor 8000 a has astationary scroll 8005 and a mobile scroll 8006 that orbit in a similarfashion to scrolls 8003 and 8004. First-stage compressor 8000 a pulls avacuum on water 8015 in first-stage evaporator 8041 causing it toevaporate and cool. The compressed vapors exiting stage one arede-superheated in packing 8020 which has water dripping over it. Thevapors entering the second-stage compressor 8000 b from second-stageevaporator 8038 are compressed and enter condenser chamber 8025 wherethey condense onto packing 8030.

Using pump 8031, the chilled water 8015 is pumped to packing 8035 whichis in countercurrent contact with house air, thus cooling the air. Thewarmed water 8036 is sucked through filter 8039 into the second-stageevaporator 8038 where some of it flashes thus cooling the water. Theflow rate is regulated by float valve 8037. Float valve 8040 regulatesthe addition of water into the first-stage evaporator 8041 where someadditional water flashes, thus cooling the water further. This chilledwater 8015 is removed by pump 8031 and contacts house air, thuscompleting the cycle.

Water 8045 from condenser 8025 is removed by pump 8032 and directed todrip over cooling tower packing 8050 which has ambient air flowingcountercurrently driven by fan 8054. As depicted in FIG. 48, the fan ispreferably driven by magnetic coupling 8055. Alternatively, it may bedriven by an independent electric motor. As the water flows through thepacking 8050, it is cooled approaching the wet-bulb temperature of theambient air. The cooled water 8051 is sucked through filter 8052 and isdrawn into condenser 8025. The water flow rate is regulated by floatvalve 8053 which directs the incoming water to drip over packing 8030.

Because both chilled water 8015 and condenser water 8045 directlycontact air, dissolved gases will be released in the vacuum of theevaporators 8038 and 8041 and condenser 8025. The noncondensible gaseswill accumulate in condenser 8025; therefore, a vacuum pump or aspiratoris needed. Accordingly, novel vacuum pump 8060 is provided.Specifically, as depicted in FIG. 48, novel vacuum pump 8060 is drivenby gear 8065 located on the main drive shaft 8007. Two reduction gears8066 and 8067 slow the rotation rate substantially. Two cammed surfaces8073 a and 8073 b are located on the slowest gear 8067. Roller 8070rides on cam 8073 b and drives piston 8071. Roller 8072 rides on cam8073 a and drives inlet valve 8075. Water 8051 is drawn into thecylinder of 8062 of vacuum pump 8060 through sprayer 8077. As piston8071 moves upward, it compresses the trapped vapors causing the watervapor to condense. The compressed noncondensible gas and excess watersprayed into the vacuum pump exit through vent 8080. To ensure bettercontact of water vapor with liquid water, packing 8078 may be placed inthe head space of vacuum pump 8060. Pump 8060 operates flooded withliquid which cools the compressor and allows water vapor to condense.Also, the water seals and lubricates the piston. Further, the water canfill dead volume allowing this pump to have an exceptional compressionratio of approximately 400:1. Although this embodiment depicts one formof vacuum pump, it is clear to one of skill in the art that one couldsubstitute the vacuum pumps depicted in other embodiments disclosedherein, including, but not limited to, pumps 10060 (FIGS. 69-70), 12060(FIGS. 80 a-b), 12402 (FIG. 81), and 12403 (FIG. 82).

Because water evaporates in evaporators 8041 and 8038 and cooling towerpacking 8050, make-up tap water 8012 and 8013 is added to the ambientair contactor (8012) and room air contactor (8013). To purge salts thatwould accumulate in the system, overflow weirs 8085 and 8086 areprovided.

The use of multiple-stage compressors as depicted in FIG. 48 providesthe following benefits:

-   -   multi-stage compression is more energy efficient than        single-stage compression;    -   an individual compressor stage is smaller than if the entire        compression were done in a single compressor; and    -   the energy efficiency is not lowered as much by mismatches        between the fixed compression ratio of the scroll compressor and        the compression ratio required by the evaporating and condensing        temperatures, which varies with the ambient temperature. By        using two stages, poorly timed vapor discharge results in less        extra work compared to a single-stage compressor.

One advantage of this embodiment is that it has no valves. Because thepressures are so low, it is not possible to use traditional check valvesthat are opened by a slight over pressure. Instead, actuated valveswould be required which creates additional mechanisms and a controlproblem. The scroll compressor eliminates the complexities associatedwith compressor valves. Multiple staging reduces the energy penaltiesassociated with compression ratio mismatches.

Still another embodiment of a scroll compressor is depicted in FIG. 49which shows a two-stage scroll compressor 8400 incorporated into cooler8801 in which the two compressors are driven by the same crank 8409. Theadvantage of this arrangement is that fewer bearings are required.

Reference numerals in FIG. 49 correspond to like elements previouslydescribed components in FIG. 48, so a further description is omitted. InFIG. 49, most of the components are analogous to those in FIG. 48;however, they are arranged slightly differently. The first-stageevaporator 8441 is concentric with the second-stage evaporator 8438.Pipes 8100 emanate radially from the second-stage evaporator 8438 andconnect to duct 8110 which directs the low-pressure vapors to the inletof the second-stage compressor. Sliding seal 8105 separates the inletsof the two compressors.

FIGS. 50-51 show an alternate and novel means of moving the mobilescroll 8004 in an orbital motion. Stationary scroll 8003 has an attachedgear 8200. (For illustration purposes, the flutes of all scrolls areremoved to reveal the internal mechanism. Also, the gear teeth areremoved to simplify the drawing.) Orbiting arm 8205 has an intermediarygear 8210 which drives gear 8220 which is attached to mobile scroll8004.

FIG. 52 shows yet another embodiment of the present invention: asingle-stage, back-to-back scroll compressor 8500 incorporated intocooler 8802. The advantage of this compressor is that the scrolldiameter can be smaller to achieve the same flow. Because of thepressure difference, the stationary scrolls must support a load. To makethem rigid requires reinforcing. Smaller diameters require lessreinforcing because there is less load, and because there is less span.Another advantage of the back-to-back scrolls is that the flutes of eachscroll can be rotated 180° with respect to each other so that the torqueis more uniform across the entire rotation.

Reference numerals on FIG. 52 correspond to previously describedcomponents so that further description is omitted. First stationaryscroll 8301 and second stationary scroll 8303 are joined by spacer 8300which provides axial, radial and angular alignment. Mobile scroll 8305has holes 8306 so that the compressed vapors in the lower chamber canescape. For simplicity, FIG. 52 shows only a single stage; however,multiple stages may be employed as well.

To reduce friction, all the scroll compressors 8000 a and 8000 b (FIG.48), 8400 (FIG. 49), and 8500 (FIG. 52) have a gap of a few thousandthsof an inch between the overlapping faces of the flutes. For instance, asdepicted in FIG. 52, gaps 8550 separate the flutes. If desired, a finemist of liquid water can be sprayed into the compressor inlet to wet thesurfaces and provide sealing as well as cooling.

Additional embodiments of the invention are directed to novellow-friction gerotor compressors 9300, 9400 and 9500, as depicted inFIGS. 53-67, useful in vapor-compression evaporative coolers as well asother applications. Unlike conventional gerotors, in which one geroterdirectly drives the other through the gerotor teeth, these low-frictiongerotors have gaps between the gerotors, and incorporate novel means tosupport and drive the gerotors. These compressors can be incorporatedinto systems such as cooler 8800 depicted in FIG. 48, or in otherembodiments such as cooler 10000 depicted in FIG. 71. These gerotorcompressors require no valves which greatly simplifies their design. Inaddition, all motion is purely rotary which is simpler to achieve thanthe orbital motion required in scroll compressors. In addition, therelative motion of the two gerotors is very slow, thus minimizing anyfriction in the wetted interior of the components. Unlike scrollcompressors, the compression ratio of the gerotor compressor, whileoperating, can be matched to the changing compression needs of the airconditioning system, thus eliminating energy waste associated with undercompressing, or over compressing, the high-pressure vapors.

One embodiment of a novel gerotor compressor having actuation means isdepicted in FIGS. 53-57. FIGS. 53 a-j depict a sequence of images as thecomponents of gerotor compressor 9300 rotate about their respectiveaxes. The inner gerotor 9302 has one less tooth than the outer gerotor9308 causing a void volume to appear between the two gerotors. Therightmost volume expands drawing low-pressure vapors into the gerotorand the leftmost volume contracts, thus expelling high-pressure vapors.Top inlet end plate 9303 and bottom outlet end plate 9305 of gerotorhousing 9301 have inlet port 9312 and outlet port 9310, respectively,that allow low-pressure vapors to enter from the top and high-pressurevapors to exit from the bottom.

FIG. 56 shows a schematic cross section of gerotor compressor 9300.Because the gerotor compressor must be large to compress the largevolumes of water vapor, friction losses and wear resulting from touchinggerotor teeth will be unacceptable; therefore it is necessary to actuatethe gerotors. The present embodiment uses novel means to actuate andsupport the gerotors. Specifically, as shown in FIG. 56, the actuationis provided by an internal gearbox 9350 that has the appropriate gearratio (i.e., in FIGS. 53-57, a 5:4 gear ratio is used). The gearbox 9350is suspended between two shafts, input shaft 9351 and output shaft 9352.Because the two shafts do not have a common center, the housing ofgearbox 9350 will not rotate as the shafts rotate. The input shaft 9351and output shaft 9352 of the gearbox 9350 rotate in the same directionbecause there are an odd number of spur gears; an idler gear 9353connects the input gear 9354 and output gear 9355.

As depicted in FIG. 57, the plate 9320 that couples the upper shaft 9351to the outer gerotor 9308 preferably has five prongs 9321 that arerecessed into the outer gerotor 9308. Because the prongs 9321 arerecessed, this allows both gerotors to be flush with the upper plate9303 of the housing 9301 which eliminates potential dead volumeassociated with the inlet port 9412.

As depicted in FIG. 56, the housing outlet port 9310 may have a fixedopening, thus fixing the compression ratio of the gerotor compressor.Alternatively, the outlet port 9310 may have a variable port mechanism.In a preferred embodiment, the discharge port has one of the variableport mechanisms depicted in FIGS. 58-60, which show three possible andnovel variable port mechanisms.

FIG. 58 depicts a variable port mechanism 9359 which has plates 9360guided by pins 9631. Springs 9632 force plates 9360 in the closed(downward) position. When actuator 9363 is slid to the right, ramp 9364forces guides 9365 to lift plates 9630 one by one, thus giving outletport 9312 a variable opening. Alternatively, rather than using actuator9363 to open plates 9360, each plate 9360 could be opened individuallyby a solenoid, or a hydraulic or pneumatic actuator.

FIG. 59 shows a variable port mechanism 9369 employing a plurality ofrigid plates 9370 that have guide loops 9371 attached to the top. Eachguide loop 9371 has a center pin 9372 and two links 9373 and 9774. Thisarrangement allows rigid plates 9730 to be connected together in amanner analogous to a bicycle chain. As slider 9375 moves leftward, itcloses outlet port 9312 and when it moves rightward, it opens outletport 9312. Spring 9376 stretches the linked plates 9370 tightly againstslider 9375.

FIG. 60 shows a variable port mechanism 9379 employing an elastomer pad9380 that has a plurality of slits 9381 that divides the pad into plates9382. As slider 9385 moves leftward, it closes outlet port 9312 and whenit moves rightward, it opens outlet port 9312. Spring 9386 stretches thepad 9380 tightly against slider 9385. Roller 9383 reduces friction ofpad 9380 against slider 9385. These mechanisms (9359, 9369 and 9379) canbe flush with end plate 9305. Also, when incorporated into laterdisclosed embodiments such as compressor 10300 depicted in FIG. 68,which has actuating spur gears 10360 and 10361, grooves may be added toplates 9360, 9370 and 9382 to accommodate the actuating gears.

Alternatively, the outlet port 9310 may have a variable port mechanismsuch as sliding mechanism 9313 depicted in FIGS. 61-62, that changes theposition of the leading edge of the discharge port, thus allowing thecompression ratio to be controlled. Sliding mechanism 9313 has a slidingcover 9314, a thin metal plate 9315 and a variable port 9316. Thesliding mechanism may be activated by servo motor 9317.

In still other embodiments, the variable port mechanism can becontrolled by the other various mechanisms disclosed herein.

Gerotor compressor 9300 may be incorporated into many types of coolers,such as cooler 8800 depicted in FIG. 48, in place of scroll compressor8000. In addition, this novel gerotor compressor could be used in anumber of applications, such as an air compressor, a compressor ofindustrial gases, a compressor for an engine (e.g., Brayton cycle), oroperated in reverse as an expander or air motor.

A primary advantage of gerotor compressor 9300 is that it has no valves.Because the pressures are so low, it is not possible to use traditionalcheck valves that are opened by a slight over pressure. Instead,actuated valves would be required which requires additional mechanismsto instantaneously open and close the valve at the precise time in thecompressor cycle, which presents a formidable control problem. Gerotorcompressor 9300 eliminates the complexities associated withinstantaneously actuated compressor valves. The variable port mechanismsshown in FIGS. 58-60 or sliding mechanism or valve 9313 shown in FIGS.61-62 can be adjusted to change the compression ratio of the gerotorcompressor, but this valve does not require instantaneous actuation;rather, it can be moved slowly (during a few seconds) to the desiredlocation. The position of this valve may be controlled by thermocouplesthat determine the evaporator and condenser temperatures. Thistemperature information would be fed to a computer that determines therequired compression ratio, and electrically actuates the sliding valveusing a servo motor 9317, a stepper motor or other means known in theart.

FIGS. 63-64 depict gerotor 9400, which is an alternate embodiment of anovel actuated low-friction gerotor compressor. In this embodiment, therelative rotation of the two gerotors is produced by two spur gears,rather than contact of the gerotor teeth. The smaller gear 9460 hasteeth on the outside diameter and the larger spur gear 9461 has teeth onthe inner diameter. The gear ratio of these two spur gears is the sameas the ratio of the number of teeth on the gerotor (in this case, 5:4).The gear set can be located at the top of the gerotors, as depicted inFIG. 63, or at the bottom. In addition to being useful in the coolingsystems disclosed herein, novel gerotor 9400 can also be used in otherapplications, such as an air compressor, a compressor of industrialgases, a compressor for an engine (e.g., Brayton cycle), or operated inreverse as an expander or air motor.

As depicted in FIG. 63, the upper shaft 9462 rotates and drives the hub9463 connected to the outer gerotor 9408. As the outer gerotor spins,the larger gear 9461 drives the smaller inner gear 9460 causing theinner gerotor 9402 to rotate. The inner gerotor spins about a fixed,nonrotating central shaft 9464. The central shaft has a crook 9465 thatestablishes the required offset of the axes of rotation for the twogerotors. If desired, a gear set 9466 can be attached to the bottom ofthe inner gerotor allowing power to be taken off for ancillaryequipment, such as the pumps.

The top stationary discharge plate 9467 with the discharge port 9468 islocated directly against the two gerotors. The discharged high-pressurevapors also must pass through perforations 9469 in the upper part of theouter gerotor and perforations 9412 in the upper plate 9403 of thehousing. There is an inlet port 9410 in the bottom of the housing 9401.

FIGS. 65-67 depict yet another low-friction embodiment, gerotorcompressor 9500, in which the outer gerotor 9508 is driven by the uppershaft 9562. The inner gerotor 9502 has rollers 9561 at the corners 9563that extend just slightly beyond the walls 9504 of the inner gerotor;thus, the rollers 9561 contact the outer gerotor 9508, but the walls9504 of the inner gerotor 9502 do not. The clearance between the wallsof the inner and outer gerotors is determined by the amount the rollers9561 extend beyond the inner gerotor wall (perhaps 0.005 inches). Theouter gerotor 9508 drives the inner gerotor 9502 through the rollercontacts. The inner gerotor 9502 is mounted on a rotary shaft 9564 thatextends out of the housing allowing auxiliaries (e.g., pumps) to bedriven from the rotating shaft. Because the relative speed of the innerand outer gerotors is relatively small (for example 300 rpm), the rollerrotation speed is not excessive (for example 2000-3000 rpm).

As noted, the present invention is directed to cooling systems havingtwo concentric chambers. The ambient air contactor is disposed in theouter chamber. The inner chamber is divided into two chambers,containing the condenser and the evaporator with the compressor inbetween. In a preferred embodiment of this cooler, the condenser is onthe bottom and the evaporator is on the top. This embodiment preferablyutilizes low-friction gerotor compressor 10300 and vacuum pump 10060incorporated into vapor-compression evaporative cooler 10000 as depictedin FIGS. 68-71. Compared to previously described coolers, the coolerdescribed in this embodiment allows for water spray to drain from thecompressor by locating the evaporator above the condenser. Water issprayed into the compressor inlet to remove superheat during compressionand provide sealing.

In addition, the novel vacuum pump 10060 disclosed herein can operate ata higher frequency because liquid water is not oscillating. The higherfrequency allows for a more compact size, and also reduces forces in thedrive train.

FIG. 68 depicts gerotor compressor 10300, and FIG. 71 shows a schematiccross section of gerotor compressor 10300 incorporated into cooler10000. Because the gerotor compressor must be large to compress thelarge volumes of water vapor, friction losses and wear resulting fromtouching gerotor teeth will be unacceptable; therefore it is necessaryto actuate the gerotors. Gerotor compressor 10300 is actuated in a novelfashion, similar to the embodiment depicted in FIG. 63-64, using a largegear 10361 with internal teeth and with a small gear 10360 with externalteeth except that the gears are located on the bottom of the compressor.

Like the gerotor compressor depicted in FIGS. 56-57, the plate 10320that couples the upper shaft 10351 to the outer gerotor 10308 has fiveprongs 10321 that are recessed into the outer gerotor 10308. Because theprongs 10321 are recessed, this allows inner gerotor 10302 to be flushwith plate 10320 and outer gerotor 10308 to be flush with housing 10301which eliminates potential dead volume. Ball bearings 10370 allow shafts10351 and 10375 to rotate within the housing 10301.

The discharge port 10310 located at the bottom of the housing may have afixed opening, thus fixing the compression ratio of the gerotorcompressor. Alternatively, the discharge port 10310 may have a variableport mechanism that changes the position of the leading edge of thedischarge port, thus allowing the compression ratio to be controlled.The variable port mechanism may be in any of the forms disclosed herein,such as the ones depicted in FIGS. 58-62, or the port may be controlledusing any other means known to one skilled in the art. In FIG. 68, servomotor 9317 is shown to indicate the possible use of the variable portmechanisms previously described. Although gerotor compressor 10300 hasbeen described in connection with the cooling systems disclosed herein,it may be used in other applications, such as an air compressor, acompressor of industrial gases, a compressor for an engine (e.g.,Brayton cycle), or operated in reverse as an expander or air motor.

FIGS. 69-70 depict a novel vacuum pump 10060 which removesnoncondensibles from the condenser. The piston 10610 is driven by acrank 10601. The piston 10610 has numerous perforations 10611 in the topthat allow vapor to flow into the cylinder 10615. A flexible flap 10612is located at the bottom of the piston 10610 which opens when the piston10610 moves upward and closes when the piston 10610 moves downward. Theopening and closing of the flap 10612 is driven both by inertia, andpressure differences across the flap 10612. During the entire operationof the vacuum pump, water 10602 is sprayed into the chamber whichcondenses water vapor as the volume reduces. Notches 10613 in the piston10610 allow liquid and compressed noncondensibles to access the checkvalve 10614 and exit the system. Although described in connection withthis particular embodiment, vacuum pump 10060 may be incorporated inplace of the vacuum pumps or aspirators of the other cooler embodimentsdescribed herein.

FIG. 71 is a schematic representation of gerotor 10300 incorporated intoa vapor-compression evaporative cooler 10000. Electric motor 10001drives the gerotor compressor 10300 through flexible coupling 10002. Thegerotor compressor pulls a vacuum on water 10015 causing it to evaporateand cool. Using pump 10031, the chilled water 10015 is pumped to packing10035 which is in countercurrent contact with house air, thus coolingthe air. The warmed water 10036 is sucked through filter 10039 into theevaporator 10038 where some of it flashes on packing 10100, thus coolingthe water; the flow rate is regulated by float valve 10037.

Cooler 10000 operates similarly to previous embodiments. Water 10045from condenser 10025 is removed by pump 10032 and directed to drip overcooling tower packing 10050 which has ambient air flowingcountercurrently driven by fan 10054. As the water flows through thepacking, it is cooled approaching the wet-bulb temperature of theambient air. The cooled water 10051 is sucked through filter 10052 andis drawn into condenser 10025. The water flow rate is regulated by floatvalve 10053 which directs the incoming water to drip over packing 10030.

Because both chilled water 10015 and condenser water 10045 directlycontact air, dissolved gases will be released in the vacuum ofevaporator 10038 and condenser 10025. The noncondensible gases willaccumulate in condenser 10025; therefore, a means for removing them,such as vacuum pump 10060, is provided. Vacuum pump 10060 is driven bygear 10065 located on the main drive shaft 10007. Using sprayers 10602,water 10051 is drawn into the vacuum pump 10060 by the vacuum. Further,as piston 10610 is driven upward by crank 10601, elastomer flap 10612opens due to its inertia. The open flap allows noncondensibles and watervapor to enter through perforations 10611. As piston 10610 is drivendownward by crank 10601, elastomer flap 10612 closes due to inertia,sealing the water vapor and noncondensibles inside. As the piston 10610compresses further, the water vapor condenses onto liquid water sprayleaving noncondensible gases and condensed water to exit through checkvalve 10614, into cooling tower 10050. Grooves or notches 10613 ensurethe check valve 10614 is not blocked when piston 10610 is fullydownward.

Because water evaporates in evaporator 10038 and cooling tower packing10050, make-up water 10012 and 10013, such as tap water, is added. Topurge salts that would accumulate in the system, overflow weirs 10085and 10086 are provided.

Like the previous gerotor embodiments, a primary advantage of gerotorcompressor 10300 is that it has no valves. In this embodiment, liquidwater will be sprayed into the compressor to eliminate superheat. Thegerotor compressor in this disclosure has low-pressure vapors enteringthe top and high-pressure vapors exiting the bottom. This arrangementallows liquid water to drain from the compressor.

FIG. 72 shows an alternative novel gerotor compressor embodiment thatcan also be used in cooler 10000 as well as the previously describedapplications. This compressor 10400 has gerotors in which the relativerotation of the two gerotors also is produced by two spur gears, thesmaller one 10460 with teeth on the outside diameter and the larger one10461 with teeth on the inner diameter. The gear ratio of these two spurgears is the same as the ratio of the number of teeth on the gerotor (inthis case, 5:4).

As depicted in FIG. 72, lower shaft 10462 rotates and drives the hub10463 connected to the inner gerotor 10402. As the inner gerotor 10402spins, the small gear 10460 drives the large gear 10461 causing theouter gerotor 10408 to rotate. The outer gerotor spins about a fixed,nonrotating central shaft 10464. The central shaft has a “crook” 10465that establishes the required offset of the axes of rotation for the twogerotors.

The top stationary inlet plate 10467 with the inlet port 10468 islocated directly against the two gerotors. The inlet low-pressure vaporsalso must pass through perforations 10469 in the connecting plate 10475to the outer gerotor 10408, and perforations 10470 in upper plate 10403of the housing. High-pressure vapors exit through discharge port 10480.The discharge port 10480 can be fixed or have a variable openingemploying the mechanisms previously described. Servo motor 9317 is shownto represent an actuation means to adjust the port opening. Shaftsupport 10481 contains ball bearings 10482 that support rotating shaft10483.

Another embodiment of the present invention is directed to alow-friction gerotor compressor without cantilevers; instead, bothgerotors are supported at two points at opposite ends of the gerotor.This gerotor can incorporate a novel means for mounting a stationaryshaft (FIGS. 74 a-d) which allows for angular and axial variation. Asdepicted in FIG. 73, gerotor compressor 11400 may be integrated intocooling systems such as cooler 11000, which is similar to cooler 10000depicted in FIG. 71.

The outer gerotor 11408 and inner gerotor 11402 of compressor 11400rotate, compress vapors, as in the previous embodiments, such as thatdepicted in FIGS. 53 a-j, so that a further description is omitted. Asindicated in FIG. 73, the gerotor is actuated by meshing a largeinternal gear 11461 and a small external gear 11460 with the same gearratio as the gerotors (in this case 4:5). As further depicted in FIG.73, gerotor compressor 11400 has a stationary central shaft 11464 withtwo crooks 11465 and 11466. One end of the stationary shaft 11464 isfixed in a pivotable mount 11470 that prevents the shaft 11464 fromrotating, but allows for angular variation. The pivotable mount couldconsist simply of a fixed rubber block 11470 (FIG. 73) with a hole inthe center to which the stationary shaft connects.

Referring again to compressor 11400 depicted in FIG. 73, the other endof stationary shaft 11464 is located in rotating bearing cup 11472. Atthe top of the compressor, there is a stationary inlet plate 11467 withan inlet hole 11468. At the bottom of the compressor there is astationary outlet plate 11477 with an outlet hole 11478 as shown in FIG.73. On either side of stationary plates 11467 and 11477 are rotatingperforated plates 11480 and 11482 that couple to the outer gerotor11408. Upper rotating perforated plate 11480 has an inlet 11484. Bottomrotating perforated plate 11482 has an outlet 11486. On either side ofthe rotating perforated plates 11480 and 11482 are perforated housingplates 11403 and 11405 that allow vapors to flow in and out through topinlet 11487 and bottom outlet 11410.

In an alternate embodiment, the housing sidewall 11430 and perforatedhousing plate 11405 can be eliminated by mounting motor 10001 on aseparate frame. Further, the rubber block 11470 can be replaced by themechanism shown in FIGS. 74 a-d.

FIGS. 74 a-d show novel pivotable mount 11490 which consists of a ring11491 with center hub 11492 connected by spokes 11493. The outer surfaceof ring 11491 is a section of a sphere which allows the ring toangularly rotate within inlet port 11487 a of upper housing plate 11403.Stationary shaft 11464 is rigidly attached to center hub 11492. Toprevent rotation of stationary shaft 11464, pin 11494 is inserted intoslot 11495 in upper housing plate 11403. Referring again to FIG. 73, afine mist of tap water may be provided through housing inlet port 11487to cool the compressor and seal rotating components. Pressure reliefvalves 11488 are preferably provided in ports 11489 of housing plate11403 to relieve excess pressure differences between the evaporator10038 and condenser 10025. Excess pressure differences could occurduring start up if the evaporator 10038 had a large quantity ofnoncondensibles (i.e., air). This unique mount allows for variations inboth axial and angular alignment while preventing shaft 11464 fromrotating. Although described in connection with the present embodiment,this mount could be used in other applications. For example, theimpellor of a centrifugal pump could be located on a shaft that ismounted to the housing using the devices illustrated in FIGS. 74 a-d.

The discharge port 11478 located on the stationary bottom plate 11477can have a fixed opening, thus fixing the compression ratio of thegerotor compressor. Alternatively, the discharge port 11478 can have avariable port mechanism such as those shown in FIGS. 58-62, that changesthe position of the leading edge of the discharge port, thus allowingthe compression ratio to be controlled. The position of the leading edgemay be set using the mechanism depicted in FIGS. 73 and 75-76. To savespace, the servo motor 11310 that sets the position of the slidingmechanism 11313 may be located outside the compressor. The servo motorrotates a threaded rod 11318 that axially positions a nonrotating nut11319 that is coupled to a bellows 11321. The bellows 11321 is filledwith a noncompressible fluid (e.g., hydraulic oil). As the bellows 11321is compressed by the servo motor 11310, the noncompressible fluid flowsdown the hollow center of the stationary shaft 11464 and extends thebellows 11320 located inside the compressor. As this bellows 11320extends, it actuates the sliding cover 9314 or 11314 (FIGS. 61-62 and75-76), actuator 9363 (FIG. 58), slider 9375 (FIG. 59) or slider 9385(FIG. 60).

Alternatively, as depicted in FIG. 77, a temperature sensor can beconstructed from liquid-containing bulb 11322. At higher temperatures,the vapor pressure of the liquid increases, causing bellows 11320 toextend and actuate the sliding discharge port cover 9314 or 11314 (FIGS.61-62 and 75-76), actuator 9363 (FIG. 58), slider 9375 (FIG. 59) orslider 9385 (FIG. 60). The movement is resisted by spring 11324, whichdetermines the functional relationship between temperature and slideposition.

FIG. 73 shows a schematic representation of gerotor compressor 11400incorporated into vapor-compression evaporative cooler 11000. Referencenumerals for elements in FIG. 73 refer to corresponding elements in FIG.71 so that a further description thereof is omitted. Cooler 11000operates in a similar fashion to cooler 10000 depicted in FIG. 71,except that electric motor 10001 directly drives the gerotor compressor.No flexible coupling is required because the pivotable mount 11470 or11490 adjusts for slight misalignments. This design has the advantagethat both gerotors are supported at each end, unlike the other designsin which one or more gerotors was cantilevered. Further, the mostprecise components (e.g., crooks 11465 and 11466) are small, so theprecision is fairly easy to achieve. In contrast, many of the otherdesigns require precise housings, which may be expensive given theirlarge size. The design shown in FIG. 73 is tolerant of imprecisionbecause of the shaft mounts 11470 or 11490 which allow for misalignment.

Like previous embodiments, another primary advantage of gerotorcompressor 11400 is that it has no valves and liquid water may besprayed into the compressor to eliminate superheat. The gerotorcompressor of this embodiment has low-pressure vapors entering the topand high-pressure vapors exiting the bottom. This arrangement allowsliquid water to drain from the compressor.

Still other embodiments of the invention are directed to integratedsystems using novel means to remove water vapor from thenoncondensibles. In these embodiments, direct contact with chilledliquid water removes much of the water vapor from the noncondensiblesstream, thus increasing the partial pressure of the noncondensibleswithout using a compressor. This innovation may be employed in asingle-stage unit suitable for the home market or it may be used inmultistage units for large buildings.

The use of chilled water to condense water vapor from thenoncondensibles was suggested in one study by the Thermal StorageApplications Research Center of the University of Wisconsin, The Use ofWater as a Refrigerant, Report No. TSARC 92-1, March 1992. However, inthis case, the author suggested the use of a metal heat exchanger withchilled water on one side and condensing water vapor on the other. Thisapproach has a severe disadvantage because of temperature differencesneeded to transfer heat. As a consequence, much of the water vapor doesnot condense because the temperature is not cold enough. In contrast,the embodiment illustrated in FIG. 78 directly contacts the water vaporwith chilled water allowing for very low temperature gradients and muchgreater water removal from the noncondensibles.

As shown in FIG. 78, removal of noncondensibles can be accomplished bypassing condenser vapor through a stripper or a packed column withchilled water flowing countercurrently. In a preferred embodiment, thecolumn contains structured packing consisting of thin sheets of PVC,such as CPVC, folded in a corrugated pattern. Alternatively, randompacking may be employed, such as ceramic saddles.

The chilled water condenses water vapor which raises the partialpressure of the noncondensibles. For example, as shown in FIG. 78,assume that at the bottom of the column, the noncondensible partialpressure is 0.04 psia and the water vapor pressure in the condenser is0.616 psia (86° F.). Further, at the top of the column, assume that thechilled water from the evaporator has a vapor pressure of 0.178 psia(50° F.). Assuming negligible pressure drop through the column, thetotal pressure is 0.656 psia at both the top and bottom of the column.Therefore, the partial pressure of noncondensibles at the top of thecolumn is 0.478 psia. At the bottom of the packed column, the ratio ofpartial pressures is $\begin{matrix}{\frac{P_{water}}{P_{noncond}} = {\frac{0.616\quad{psia}}{0.040\quad{psia}} = {15.4 = \frac{15.4\quad{lb}\quad{mol}\quad{water}}{{lb}\quad{mol}\quad{noncondensibles}}}}} & (13)\end{matrix}$At the top of the packed column, the ratio of partial pressures is$\begin{matrix}{\frac{P_{water}}{P_{noncond}} = {\frac{0.178\quad{psia}}{0.478\quad{psia}} = {0.372 = \frac{0.372\quad{lb}\quad{mol}\quad{water}}{{lb}\quad{mol}\quad{noncondensibles}}}}} & (14)\end{matrix}$

Thus, using this very simple device, the noncondensible pressure ratioincreases by a factor of twelve while simultaneously removing almost 98%of the water vapor. Assuming the packing is able to operate nearequilibrium, the required amount of chilled water is $\begin{matrix}\begin{matrix}{\hat{m} = {( {\frac{15.4\quad{lb}\quad{mol}\quad{water}}{{lb}\quad{mol}\quad{noncondensibles}} - \frac{0.372\quad{lb}\quad{mol}\quad{water}}{{lb}\quad{mol}\quad{noncondensibles}}} ) \times}} \\{\frac{{lb}\quad{mol}\quad{noncondensibles}}{29\quad{lb}} \times \frac{18\quad{lb}\quad{water}}{{lb}\quad{mol}\quad{water}} \times} \\{\frac{1000\quad{Btu}}{{lb}\quad{water}} \times \frac{{lb}\quad{chilled}{\quad\quad}{{water} \cdot {{^\circ}F}}}{1\quad{Btu}} \times \frac{1}{( {86 - 50} ){{^\circ}F}}} \\{= \frac{259\quad{lb}\quad{chilled}\quad{water}}{{lb}\quad{noncondensibles}}}\end{matrix} & (15)\end{matrix}$

Based upon the solubility of air in both the chilled water and condenserwater, the mass flow rate of noncondensibles is about 0.051 lb/h for a1-ton (12,000 Btu/h) air conditioner. Therefore, the chilled water flowrate for the water stripper in a 1-ton air conditioner is$\begin{matrix}\begin{matrix}{{\overset{.}{m}}_{stripper} = {259{\frac{{lb}\quad{chilled}\quad{water}}{{lb}\quad{noncondensibles}} \times \frac{0.051\quad{lb}\quad{noncondensibles}}{h}}}} \\{= {13.2\frac{{lb}\quad{chilled}{\quad\quad}{water}}{h}}}\end{matrix} & (16)\end{matrix}$

Optimization studies (see FIG. 3) show that the best temperature changefor the chilled water that circulates through the house is 4° C. (7°F.); therefore, the required water flow rate for a 1-ton air conditioneris $\begin{matrix}{{\overset{.}{m}}_{house} = {{12\text{,}000{\frac{Btu}{h} \times \frac{{lb}\quad{chilled}{\quad\quad}{{water} \cdot {{^\circ}F}}}{1\quad{Btu}} \times \frac{1}{7{{^\circ}F}}}}\quad = {1714\frac{{lb}\quad{chilled}{\quad\quad}{water}}{h}}}} & (17)\end{matrix}$

Thus, the chilled water flowing to the stripper is only about 0.8% ofthe chilled water circulating through the house, which has almost anegligible effect on the compressor power requirements.

A variety of coolers may use chilled water to remove noncondensiblesincluding the system depicted in FIG. 73. For instance, FIG. 79 depictsa single-stage evaporator cooling system 12000 that uses a gerotorcompressor such as gerotor compressor 11400 depicted in FIG. 73. Coolingsystem 12000 is similar to cooling system 11000 in FIG. 73, except thata gerotor vacuum pump 12060 is used in place of vacuum pump 10060 toremove noncondensibles.

FIGS. 79-81 show a gerotor vacuum pump 12060 which operates similarly tothe main compressor; however, it is much smaller. For example, the maincompressor of a 1-ton air conditioner has a volumetric flow of about 470ft³/min whereas the vacuum pump must process only 0.24 ft³/min for anidentical air conditioner. The center gerotor 12003 is mounted on thelower portion of the main drive shaft 12004 whereas the outer gerotor12005 is positioned by guide rollers 12006. Alternatively, the outergerotor may be mounted within a single large ball bearing. A gear on thecenter gerotor 12003 can drive a gear on the outer gerotor 12005—as withthe main compressor—or the inner gerotor can drive the outer gerotordirectly without an intervening gear. Because the compression ratio isquite high (about 22:1), the temperature rise of the exhaust gas couldbe quite dramatic; therefore, it is beneficial to introduce liquid water12220 into the vacuum pump. The optimal location is to introduce theliquid water immediately after the intake portion of the cycle. Enoughliquid water can be introduced to fill void volumes in the gerotor thusallowing very high compression ratios to be achieved. Bothatmospheric-pressure air and liquid water will be discharged from theair conditioner. A check valve 12230 is preferably located in thedischarge line to prevent leakage of atmospheric air into the airconditioner. Optionally, an accumulator can be placed between thegerotor vacuum pump and the check valve so the check valve does not haveto cycle rapidly.

FIG. 79 depicts a schematic of the gerotor compressor 11400 and gerotorvacuum pump 12060 incorporated into a single-stage vapor-compressionevaporative cooler 12000. Electric motor 12001 directly drives gerotorcompressor 11400. No flexible coupling is required because the pivotablemount 11470 adjusts for slight misalignment. The gerotor compressor11400 pulls a vacuum on water 12015 causing it to evaporate and cool.Using pump 12031, the chilled water 12015 is pumped to packing 12035which is in countercurrent direct contact with house air, thus coolingthe air. The warmed water 12036 is sucked through filter 12039 into theevaporator 12038 where some of it flashes on packing 12100 thus coolingthe water; the flow rate is regulated by float valve 12037.

As shown in FIG. 79, bellows 12300 actuates a variable discharge port12011. In a preferred embodiment, motor 12310, which is preferably aservo motor, drives a nonrotating nut 12320 which actuates bellows 12331which, in turn, actuates bellows 12300 which adjusts variable dischargeport 12011. Alternatively, the liquid-containing bulb system shown inFIG. 77 could actuate bellows 12300.

Water 12045 from condenser 12025 is removed by pump 12032 and directedto drip over ambient air contactor packing 12050 which has ambient airflowing countercurrently driven by fan 12054. As the water flows throughthe packing, it is cooled approaching the wet-bulb temperature of theambient air. The cooled water 12051 is sucked through filter 12052 andis drawn into condenser 12025. The water flow rate is regulated by floatvalve 12053 which directs the incoming water to drip over packing 12030.

Because both chilled water 12015 and condenser water 12045 directlycontact air, dissolved gases will be released in the vacuum ofevaporator 12038 and condenser 12025. The noncondensible gases willaccumulate in condenser 12025; therefore, a vacuum pump or aspirator isprovided. As indicated in the embodiment shown in FIG. 79, gerotorvacuum pump 12060 is preferably used. Chilled water 12210 flows overpacking 12200 to remove noncondensibles. In a preferred embodiment thisis a structured packing consisting of corrugated PVC sheet.Alternatively, it could be a random packing of ceramic saddles. Somewater 12220 may be sprayed into gerotor vacuum pump 12060 for coolingand sealing purposes. Additionally, the volume of water will besufficiently large to fill voids in the gerotor allowing very highcompression ratios to be achieved. Discharge from the gerotor vacuumpump 12060 is directed through check valve 12230 and ultimately is sentto the ambient air contactor 12050.

Because water evaporates in evaporator 12038 and ambient air contactor12050, make-up water, such as tap water 12012 and 12013, is added. Topurge salts that would accumulate in the system, overflow weirs 12085and 12086 are provided.

Liquid water is preferably sprayed into the compressor 11400 toeliminate superheat. The gerotor compressor in this disclosure haslow-pressure vapors entering the top and high-pressure vapors exitingthe bottom. This arrangement allows liquid water to drain from thecompressor. The gerotor in this embodiment has no cantilevers allowingfor more reliable operation.

The method used to purge noncondensibles from the system allows for theuse of an aspirator, or a small vacuum pump, such as pump 12060 and theothers disclosed herein, because the water vapor has been largelyremoved. Also, because it is not necessary to condense water vaporsinside the vacuum pump, it can be operated at high speed whicheliminates the need for gear reduction, a potential maintenance problem.

In addition to vacuum pump 12060, this invention is also directed toother simple, efficient and novel vacuum pumps that can removenoncondensibles from a vapor-compression evaporative cooler or be usedin other applications requiring a vacuum pump. These novel vacuum pumpshave the unusual ability to isothermally compress a mixture ofnoncondensible and condensible gases to a very high compression ratio.

FIG. 81 is a schematic depicting a preferred embodiment of this vacuumpump. As depicted in FIG. 81, vacuum pump 12402 consists of two columns12404 a and 12404 b which are partially filled with liquid and arejoined by lower connecting chamber 12405. The two columns have checkvalves 12414 a and 12414 b and inlet valves 12415 a and 12415 b at theirupper ends. The liquid in each column is driven to oscillate by piston12406 located in lower connecting chamber 12405, between the bottom ofcolumns 12404 a and 12404 b. In FIG. 81, piston 12406 is magnetic and isdriven by a magnetic field induced by electric coil 12408. Springs 12410a and 12410 b at each end 12407 a and 12407 b of lower connectingchamber 12405 act as stops. When applied to the vapor-compressionevaporative cooler, the preferred liquid is water. However, for othervacuum pump applications, nonvolatile liquids (e.g., vacuum pump oil)could by employed to achieve high vacuums.

The gas and vapor mixture is introduced into center connecting inletduct 12412. A solenoid 12413 opens one inlet valve 12415 a and closesthe other inlet valve 12415 b so the gas/vapor is directed to thecolumn, i.e., 12404 a, in which the liquid is moving downward. Water isconstantly sprayed by sprayers 12417 a and 12417 b into each watercolumn, creating water sprays 12419 a and 12419 b. In the water columnthat is rising (in this example, column 12404 b), the water vaporscondense on the water spray 12419 b and the noncondensible gas becomescompressed. When the water gets to the top of the column, the respectivecheck valve, 12414 b, opens and releases the excess water and compressednoncondensible gas.

Another embodiment of a novel vacuum pump useful in removingnoncondensibles or in other applications is depicted in FIG. 82. In thisembodiment, like reference numerals refer to similar elements describedin the previous embodiment so that a further description thereof isomitted. As indicated in FIG. 82, vacuum pump 12403 is similar inconfiguration and operation to the previous compressor 12402 with theexception of the piston and the lower connecting chamber. Piston 12422is disposed in lower connecting chamber 12423 and is connected to pistonsolenoid 12420 by a rod 12424. In compressor 12403, piston solenoid12420 is in a fixed position and drives piston 12422 back and forth.Alternatively, piston 12422 could be coupled to a reversing motor by athreaded rod. In this third and also novel embodiment, piston 12422 isdriven back and forth as the motor reverses direction.

The coolers described above are relatively simple and suitable for homeuse. To reduce high utility bills, large air conditioning installationscan use even more complex systems to achieve greater energy efficiency.Accordingly, this invention is also directed to highly efficientmultistage coolers that employ the novel feature of multistagecondensation, in addition to multistage evaporation. Although multistagethrottling is known with conventional air conditioning systems,multistage evaporators are rarely used. Compressors useful in multistagecoolers include both the novel positive displacement, low-frictioncompressors previously described as well as conventional dynamiccompressors. These coolers also utilize novel means to removenoncondensibles.

FIG. 83 depicts one such energy-efficient system which employs multipleevaporator stages. Compared to a single-stage system in which all of thewater is evaporated at the lowest pressure, the multistage evaporator invapor-compression evaporative cooler 13000 allows some of the water tobe evaporated at higher pressures which reduces compression energy.

Referring to FIG. 83, in vapor compression evaporative cooler 13000,chilled water countercurrently directly contacts air from the buildingin room air contactor 13110. Because the water is cold, it both coolsthe air and condenses moisture out of the air. The warmed water from theroom-air contactor 13110 flows countercurrently through a series ofevaporators 13120. Water vaporizes in each evaporator making the liquidcolder in each successive stage. Once the water is fully chilled, it isreturned to the room air contactor 13110 via a cold pump 13121.

Multiple compressors 13130 are used so that vapors can be drawn off ofeach evaporator. To reduce the superheating of the water, liquid may beinjected directly into the compressors 13130 as described in earlierembodiments. Centrifugal or axial compressors generally are not tolerantof liquid droplets, so intercooling can be accomplished by sprayingliquid 13125 into the vapor space of the evaporator 13120. In this case,the evaporator chamber serves two purposes: it is an evaporator and ade-superheater. The source of the liquid may be tap water or coolingtower water, whichever is colder.

The vapor discharge from the last compressor 13130 is directed to acondenser 13160 where it contacts water that is near the wet-bulbtemperature of the ambient environment. As the vapors condense on thewater, the temperature rises. This hot water is pumped out of thecondenser via pump 13161 into ambient-air contactor 13150 (coolingtower).

Because water is evaporating both in the evaporators and ambient-aircontactor, make-up water 13154 and 13155 is provided. Ordinary tap wateris envisioned as the make-up water. Treated water should not benecessary because there are no heat exchange surfaces. Salts will buildup due to the evaporation, so salt water purges 13116 and 13117 areprovided.

Noncondensible gases are dissolved into the water in both the room-aircontactor 13110 and ambient-air contactor 13150. These gases arereleased in the low-pressure evaporators 13120 and condenser 13160,respectively. A small compressor train 13170 will draw vapors from thecondenser 13160 to remove the noncondensible gases. Intercooling isprovided by spraying water 13140 from the ambient-air contactor 13150between the compressor stages to condense the water vapors. Thenoncondensible partial pressure rises in each stage until it reaches 1atm and can be discharged directly to the ambient air.

The following analysis describes the energy efficiency of the systemdepending upon the number of stages employed.

One Stage

The compressor work W_(comp) per unit of heat absorbed in the evaporatorQ_(evap) is the inverse of the coefficient of performance (COP)$\begin{matrix}{\frac{W_{comp}}{Q_{evap}} = {\frac{1}{COP}\quad = {{\frac{T_{4} - T_{1}}{T_{1}}\frac{1}{\eta_{motor}}\frac{1}{\eta_{comp}}\frac{1}{\eta_{cycle}}}\quad = {\frac{( {T_{5} + {\Delta\quad T_{hot}}} ) - T_{1}}{T_{1}}\frac{1}{\eta_{motor}}\frac{1}{\eta_{comp}}\frac{1}{\eta_{cycle}}}}}} & (18)\end{matrix}$where temperatures are defined in FIG. 83 and

W_(comp)=compressor work

Q_(evap)=total heat absorbed in evaporator

COP=coefficient of performance

T₄=water temperature exiting condenser

T₁=water temperature exiting coldest evaporator

η_(motor)=motor efficiency

η_(comp)=compressor efficiency

η_(cycle)=thermodynamic efficiency of the cycle relative to Carnot

ΔT_(hot)=T₄−T₄=temperature differences between water exiting thecondenser and water exiting the ambient air contactor

The work for the cold pump is $\begin{matrix}{W_{cold} = {{V\quad\Delta\quad P\frac{1}{\eta_{pump}}}\quad = {\frac{Q_{evap}}{\rho\quad C_{p}\Delta\quad T_{cold}}\Delta\quad P\frac{1}{\eta_{pump}}}}} & (19)\end{matrix}$ $\begin{matrix}{\frac{W_{cold}}{Q_{evap}} = {\frac{\Delta\quad P}{\rho\quad C_{p}\Delta\quad T_{cold}}\frac{1}{\eta_{pump}}}} & (20)\end{matrix}$where

W_(cold)=work for the cold-water pump

V=volumetric flow rate of water through the cold pump

ΔP=pressure differences generated by the pump

η_(pump)=pump efficiency

ρ=water density

C_(p)=water heat capacity

ΔT_(cold)=T₆−T₁=temperature difference between water exiting theroom-air contactor and the water exiting the coldest evaporatorIf a turbine is employed to reduce the pumping energy requirements, thework for the cold pump is $\begin{matrix}{\frac{W_{cold}}{Q_{evap}} = {\frac{\Delta\quad P}{\rho\quad C_{p}\Delta\quad T_{cold}}( {\frac{1}{\eta_{pump}} - \eta_{turbine}} )}} & (21)\end{matrix}$where

η_(turbine)=turbine efficiencyThe work for the hot pump is $\begin{matrix}{\frac{W_{hot}}{Q_{evap}} = {\frac{\Delta\quad P}{\rho\quad C_{p}\Delta\quad T_{hot}}( {1 + \frac{1}{COP}} )\frac{1}{\eta_{pump}}}} & (22)\end{matrix}$where $\frac{1}{COP}$is from Equation 18. If a turbine is employed to reduce the pumpingenergy requirements, the work for the hot pump is $\begin{matrix}{\frac{W_{hot}}{Q_{evap}} = {\frac{\Delta\quad P}{\rho\quad C_{p}\Delta\quad T_{hot}}( {1 + \frac{1}{COP}} )( {\frac{1}{\eta_{pump}} - \eta_{turbine}} )}} & (23)\end{matrix}$The total work is $\begin{matrix}{\frac{W_{tot}}{Q_{evap}} = {\frac{W_{comp}}{Q_{evap}} + \frac{W_{cold}}{Q_{evap}} + \frac{W_{hot}}{Q_{evap}}}} & (24)\end{matrix}$Two Stage

In the case of a two-stage compressor, assuming half the load is takenby each stage, the compression work is $\begin{matrix}{{W_{comp} = {{\frac{1}{2}{Q_{evap}( \frac{1}{{COP}_{1}} )}} + {\frac{1}{2}{Q_{evap}( \frac{1}{{COP}_{2}} )}}}}{\frac{W_{comp}}{Q_{evap}} = {\frac{1}{2}( {\frac{1}{{COP}_{1}} + \frac{1}{{COP}_{2}}} )}}} & (25) \\\begin{matrix}{\frac{W_{comp}}{Q_{comp}} = \frac{1}{COP}} \\{= {\frac{1}{2}\begin{bmatrix}{\frac{( {T_{5} + {\Delta\quad T_{hot}}} ) - T_{1}}{T_{1}} +} \\\frac{( {T_{5} + {\Delta\quad T_{hot}}} ) - ( {T_{1} + {{1/2}\Delta\quad T_{cold}}} )}{( {T_{1} + {{1/2}\Delta\quad T_{cold}}} )}\end{bmatrix}}} \\{\frac{1}{\eta_{motor}}\frac{1}{\eta_{comp}}\frac{1}{\eta_{cycle}}}\end{matrix} & (26)\end{matrix}$When determining the total work using Equation 24, Equation 26 is usedto calculate the compressor work. The cold pump Equation 21 will be thesame. The hot pump Equation 23 is the same, except that Equation 26 isused for $\frac{1}{COP}.$Three Stage

In the case of a three-stage compressor (as illustrated in FIG. 83), thecompressor work is $\begin{matrix}\begin{matrix}{\frac{W_{comp}}{Q_{evap}} = \frac{1}{COP}} \\{= {\frac{1}{\eta_{motor}}\frac{1}{\eta_{comp}}\frac{1}{\eta_{cycle}}\frac{1}{3}}} \\{\begin{bmatrix}\begin{matrix}{\frac{( {T_{5} + {\Delta\quad T_{hot}}} ) - T_{1}}{T_{1}} +} \\{\frac{( {T_{5} + {\Delta\quad T_{hot}}} ) - ( {T_{1} + {{1/3}\Delta\quad T_{cold}}} )}{( {T_{1} + {{1/3}\Delta\quad T_{cold}}} )} +}\end{matrix} \\\frac{( {T_{5} + {\Delta\quad T_{hot}}} ) - ( {T_{1} + {{2/3}\Delta\quad T_{cold}}} )}{( {T_{1} + {{2/3}\Delta\quad T_{cold}}} )}\end{bmatrix}}\end{matrix} & (27)\end{matrix}$All other equations and procedures are the same.n Stage

One can generalize to an n-stage compressor as follows $\begin{matrix}\begin{matrix}{\frac{W_{comp}}{Q_{evap}} = \frac{1}{COP}} \\{= {\frac{1}{\eta_{motor}}\frac{1}{\eta_{comp}}\frac{1}{\eta_{cycle}}\frac{1}{n}}} \\{\sum\limits_{i = 0}^{n - 1}\lbrack \frac{( {T_{5} + {\Delta\quad T_{hot}}} ) - ( {T_{1} + {\frac{i}{3}\Delta\quad T_{cold}}} )}{( {T_{1} + {\frac{i}{3}\Delta\quad T_{cold}}} )} \rbrack}\end{matrix} & (28)\end{matrix}$Analysis

This system was analyzed using the following assumptions

-   -   η_(motor)=0.9 (high efficiency due to large scale)    -   η_(comp)=0.8 (high efficiency due to large scale)    -   η_(cycle)=0.97 (from FIG. 21, Reducing Energy Costs in        Vapor-Compression Refrigeration and Air Conditioning Using        Liquid Recycle—Part II: Performance, Mark Holtzapple, ASHRAE        Transactions, Vol. 95, Part 1, 187-198 (1989))    -   η_(pump)=η_(turbine)=0.5    -   T₁=285.4 K=12° C.=54° F.    -   ΔT_(cold)=11 K=20 F.°    -   ΔT_(hot)=4 K=7 F.°    -   ρ=1000 kg/m³    -   C_(P)=4189 J/(kg·K)    -   ΔP=101, 330 Pa=1 atm

FIG. 84 shows the results of the analysis with no turbines and FIG. 85shows the results with turbines. For comparison purposes, astate-of-the-art water chiller is available from Trane (CFCs: TodayThere Are Answers, FIG. 18, CFC-ARTICLE-1, The Trane Company, 3600Pammel Creek Rd., La Crosse, Wis., 54601-7599) that requires 0.50 kW/tonat standard ARI conditions (cold side=44° F. leaving evaporator, 54° F.entering evaporator; hot side=85° F. entering condenser (our T₅), 95° F.leaving condenser; cooling tower=7° F. approach temperature, 78° F. wetbulb temperature). According to FIG. 84 (no turbines, three stages), thedisclosed system with T₅=85° F. (29.4° C.) requires only 0.37 kW/ton.According to FIG. 85 (with turbines, three stages), the disclosed systemwith T₅=85° F. (29.4° C.) requires only 0.35 kW/ton. Thus, the energyrequirement of the multistage vapor-compression evaporative cooler isapproximately 70% of the current state-of-the-art system.

An important consideration is the effect of noncondensibles on thesystem. A 1-ton unit has a noncondensible load of about 0.0023 lbmole/h.If its partial pressure in the condenser is 0.05 psia, then thetheoretical work requirement (assuming isothermal compression) is$\begin{matrix}{\begin{matrix}{\frac{W_{purge}}{Q_{evap}} = {{nRT}\quad\ln\quad\frac{P_{2}}{P_{1}}}} \\{= {( \frac{0.0023\quad{lbmole}}{{ton} \cdot h} )( \frac{1.986\quad{Btu}}{{lbmole} \cdot {\,^{{^\circ}}R}} )( \frac{{kW} \cdot h}{3413\quad{Btu}} )}}\end{matrix}\begin{matrix}{{( {460 + 85} ){\,^{{^\circ}}R}\quad\ln\frac{14.7\quad{psia}}{0.05\quad{psia}}} = {0.00414\frac{kW}{ton}}} \\{= {4.14\frac{W}{ton}}}\end{matrix}} & (29)\end{matrix}$where

W_(purge)=compressor work required to purge noncondensibles

n=moles of noncondensibles to be purged

R=universal gas constant

P₂=final discharge pressure (ambient pressure)

P₁=intake partial pressure of noncondensible gas

If the partial pressure of noncondensibles in the condenser is reducedto 0.01 psia, then the work requirement increases to 5.32 W/ton.Assuming the compressor is 50% efficient, then the work requirement forpurging noncondensibles is only about 10 W/ton, which falls within the“noise.”

For a very large chiller (300 to 2500 ton), the compressor is likely tobe centrifugal. To get a sense of the scale, the low-pressure compressorwill be designed assuming a 3-compressor 500-ton unit. The low-pressurecompressor may actually have a number of stages within it. The requiredhead per stage is $\begin{matrix}{H = {\frac{1\text{,}545}{M_{w}}\frac{k}{k - 1}{T_{1}\lbrack {r^{{({k - 1})}/k} - 1} \rbrack}}} & (30)\end{matrix}$where

H=head, ft·lb_(f)/lb_(m)

M_(w)=molecular weight=18 lb_(m)/lbmole

r=compression ratio, dimensionless

k=1.323 for water

T₁=inlet temperature=54° F.=514° R.

The compression ratio of each stage in the low-pressure compressor canbe calculated from $\begin{matrix}{r = ( \frac{P_{2}}{P_{1}} )^{1/n}} & (31)\end{matrix}$where

P₂=discharge pressure of low-pressure compressor=0.311 psia (assumed)

P₁=inlet pressure of low-pressure compressor=0.202 psia (assumed)

n=number of stages within the low-pressure compressorThe discharge pressure, P₂, was calculated as $\begin{matrix}\begin{matrix}{P_{2} = {0.202\quad{{psia}( \frac{\quad{0.744\quad{psia}}\quad}{0.202\quad{psia}} )}^{1/3}}} \\{= {0.311\quad{psia}}}\end{matrix} & (32)\end{matrix}$where 0.744 psia is the pressure of a 92° F. condenser.

FIG. 86 shows a generalized compressor chart indicating regions wherepiston, centrifugal, axial, and drag compressors are appropriate. Thegeneralized correlation for a single stage within the low-pressurecompressor is made in terms of specific speed, N_(s), and specificdiameter, D_(s), defined as follows $\begin{matrix}{N_{S} = \frac{N\sqrt{Q}}{H^{3/4}}} & (33) \\{D_{S} = \frac{{DH}^{1/4}}{\sqrt{Q}}} & (34)\end{matrix}$where

N=rotational speed, rpm

Q=inlet volumetric flow rate, ft³/s

H=head, ft·lb_(f)/lb_(m)

D=diameter, ftThe volumetric flow at the inlet to the low-pressure compressor for athree-compressor, 500-ton system is $\begin{matrix}\begin{matrix}{Q = {\frac{1}{3} \times 500\quad{ton} \times \frac{12\text{,}000\quad{Btu}}{{ton} \cdot h} \times \frac{{lb}_{m}{water}}{1065\quad{Btu}} \times}} \\{\frac{h}{3600\quad s} \times \frac{1517\quad{ft}^{3}}{{lb}_{m}{water}}} \\{= {791\quad{ft}^{3/S}}}\end{matrix} & (35)\end{matrix}$The factor ⅓ results because the low-pressure compressor takes only athird of the load.

FIG. 86 shows that a centrifugal compressor with N_(s)=60 and D_(s)=2 isabout 80% efficient. Using Equations 33 and 34, the corresponding speedand diameter can be calculated. $\begin{matrix}{N = \frac{N_{S}H^{3/4}}{\sqrt{Q}}} & (36) \\{D = \frac{D_{S}\sqrt{Q}}{H^{1/4}}} & (37)\end{matrix}$

The tip speed, v, is $\begin{matrix}{v = \frac{\pi\quad{DN}}{60}} & (38)\end{matrix}$where v is in ft/s.

FIG. 87 shows the results of the compressor analysis. The tip speed fora single stage is acceptable as is the rotational speed, so a singlestage should be sufficient for the low-pressure compressor.

Advantages of this system include efficiency. In addition, waterchemistry is not so important because there are no heat exchangesurfaces.

FIG. 88 depicts a schematic of another multistage cooler, multistageevaporative cooler 13100. This cooler is similar to cooler 13000 exceptthat packed columns 13180 are used to eliminate superheat. Referencenumerals in FIG. 88 refer to corresponding elements in FIG. 83, so thata further description thereof is omitted.

Referring now to FIG. 88, in order to eliminate superheat after eachcompression stage, the vapors exiting the compressors 13130 are passedcountercurrently through a packed column 13190 with liquid passingdownward. Although centrifugal compressors may be employed in theselarge-scale systems, it is also possible to use large gerotorcompressors. The noncondensibles that accumulate in the condenser arepassed countercurrently through the packed column 13180 with chilledwater flowing downward in direct contact with the water vapor, whichcondenses most of the water vapor, as described before. In a preferredembodiment, structured packing of corrugated PVC sheets as describedearlier is used. The noncondensibles are then removed by a vacuum pump,shown here as a multistage compressor train 13170.

An even more efficient air conditioning system, vapor compressionevaporative cooler 13200 is shown in FIG. 89. Reference numerals in FIG.89 refer to corresponding elements in FIG. 83, so that a furtherdescription is omitted. In cooler 13200, both multistage evaporators13120 and multistage condensers 13160 are used. In addition, multiplepacked columns 13180 are used. The cooling water in the condensers 13160flows countercurrently to the water in the evaporators 13120. Thisminimizes the pressure difference between the evaporators and thecondensers, thus promoting the greatest energy efficiency. To preventsuperheating in each compressor and to promote energy efficiency, liquidwater 13135 may be sprayed directly into the compressor 13130. Becausehigh-speed centrifugal compressors can be damaged by liquid water, it ispreferred to employ gerotor compressors.

The preferred embodiments disclosed herein include a number of novelcooling systems that use water as the working fluid, a number of novelpositive displacement and low-friction compressors that are useful inthe disclosed coolers and other applications, and a number of novelmeans for removing noncondensibles, however, the specific embodimentsand features disclosed herein are provided by way of example only andare not intended as limitations on the scope of the invention. As willbe clear to one of skill in the art, each of the various compressors maybe adapted for use in the different disclosed cooling systems as well asother applications, and are in no way limited to the specific coolingsystem in which they are depicted. In addition, as is clear to one ofskill in the art, the variable port mechanisms, seals, mounting systemsand other novel components of the different compressors disclosed hereincan be easily interchanged by one of skill in the art, as can thedifferent novel vacuum pumps and compressors useful in removingnoncondensibles. In addition, means for inhibiting microorganisms, suchas an ozone generator, can be incorporated into any of the disclosedsystems. It will be easily understood by those of ordinary skill in theart that variations and modifications of each of the disclosedembodiments can be easily made within the scope of this invention asdefined by the following claims.

1-90. (canceled)
 91. A vacuum pump comprising a cylinder, a pistondisposed in said cylinder, an inlet valve disposed in said cylinder, asprayer that draws water into said cylinder, and a vent disposed in saidcylinder for discharging noncondensibles and excess water, and whereinsaid vacuum pump is driven by a gear mounted on a main drive shaft, saidgear connected to a plurality of reduction gears, wherein a first camsurface and a second cam surface are mounted on one of said reductiongears, a first roller rides on said first cam surface and a secondroller rides on said second cam surface, and said first roller drivessaid piston and said second roller drives said inlet valve.
 92. A vacuumpump comprising a cylinder, a piston disposed in said cylinder, a crank,a check valve disposed in said cylinder, and means for spraying waterinto said cylinder of said vacuum pump, wherein said piston is driven bysaid crank in a first and a second direction, said piston comprising afirst end, a second end, a plurality of notches, a plurality ofperforations extending from said first end to said second end, and aflexible flap attached to said second end of said piston and coveringone or more of said perforations, wherein said flap opens when saidpiston moves in said first direction and closes when said piston movesin said second direction.
 93. A vacuum pump comprising: a first columnand a second column, said columns being partially filled with liquid andhaving a vapor space; means for causing said liquid to oscillate in saidcolumns; inlet means for allowing uncompressed gas to enter each of saidcolumns; outlet means for discharging compressed gas from each of saidcolumns; and means for spraying a fine mist of liquid into said vaporspace of said first and said second columns.
 94. The pump of claim 93,wherein said means for causing oscillation comprises a chamberconnecting said first and said second columns and a reciprocating pistondisposed in said chamber.
 95. The pump of claim 93, wherein said outletmeans comprises a check valve.
 96. A gerotor vacuum pump comprising anouter gerotor and a center gerotor disposed within said outer gerotor,wherein said center gerotor is mounted on a main drive shaft and saidouter gerotor is positioned by a plurality of guide rollers.
 97. Agerotor vacuum pump comprising an outer gerotor and a center gerotordisposed within said outer gerotor, wherein said center gerotor ismounted on a main drive shaft and said outer gerotor is mounted within asingle ball bearing.
 98. A pivotable mounting apparatus for mounting astationary shaft to a housing, which prevents rotation of said shaft,but allows for angular and axial variation, comprising a ring coupled tosaid shaft having a spherical outer diameter disposed within acylindrical shaped opening in said housing.
 99. A low-friction rotaryshaft seal comprising: a journal for receiving a rotary shaft, saidjournal configured to create a gap between said shaft and said journal,said journal further comprising a journal face; means for supplyingwater to said gap; and a bellows seal, said seal resting on said journalface when said shaft is stationary and lifting off said face when saidshaft rotates.
 100. A method for removing water vapor fromnoncondensibles in a stream of air and water vapor comprising the stepof passing the stream through a packed column with chilled water flowingtherein concurrently, wherein the packed column comprises corrugatedpolyvinyl chloride.